:|/,,.^':';^:i;-'i^p^^ 



■m/mmm^mi 




Class. 
Book._ 



Coipghtl^?_ 






'-'^ 



caazmiGHT DEPosm 






^•.«>;*>-Xi> V 




.■^' \ 



X 



r^. 





NEW YORK BELTING & PACKING CO. 

124-126 WEST LAKE STREET 

CHICAGO 



1 




NEW YORK BELTING & PACKING CO. 

218-220 CHESTNUT STREET 

ST. LOI.TIS 



^ 





Hliwj m 






' I'V 


'' w 




i ^ 




^^& '^^ 


1 i ^ 




''ji 

np 




Nl 


" ■ 1 






V 
K 


ni:j 


, ,_| 


p 








" 












i 


1 

M 




i 


fi 


■ 


m 


m 


?'" 




■M] 


.1 i i 


\ 


n 
1 




1 ;i "'■;! 

■ 


1 












^K^ 




Ir -1*^15' '1 


1 'd 


k 


flUUMWift'- ' JI 


'\ . !■■. J 



Ins: 



h\ 



A MANUAL 
STEAM ENGINEERING 

COMPRISING 

Instructions, Suggestions and Illustrations 

FOR Progressive Steam Engineers, 

Concerning the Application to 

Modern Daily Practice of 

THE Approved Theory of 

Steam Engineering 

W. H. WAKEMAN 

AUTHOR OF 

"engineering practice and theory" 
"practical guide for firemen" 

AND numerous ARTICLES FOR 
THE MECHANICAL PRESS 



New York Belting & Packing Company 

91-93 Chambers St., New York 

124-126 West Lake St., Chicago 
218-220 Chestnut St., St. Louis 
821-823 Arch St., Philadelphia 
519 Mission St., San Francisco 
65 Pearl St., Boston 
420 First A^e., Pittsburgh 






^' 



\ 



Copyright, 1920 



New York Belting anp Packing 
Company 



^ 



i 






St? -B "^^ 

©CI,A576279 



f 



IL PREFACE 

Ca) 
' I "^HE object sought in presenting this 
volume to steam engineers and 
others interested in steam engineering 
is to provide a convenient and reliable 
reference book, containing data wanted 
in every-day practice, arranged in con- 
venient form, with sufficient explana- 
tion to render the matter both interest- 
ing and instructive. 

It contains many tables, which are 
arranged for use in connection with the 
reading matter on the same subject, 
thus condensing information into con- 
venient space, and rendering it available 
for the busy workers. That the work 
will prove satisfactory to readers for 
whom it is intended is the sincere wish of 

THE AUTHOR 



DEDICATION 

'npO the intelligent men who design, 
construct and operate the steam 
machinery of this country, keeping it 
in repair imder a great variety of con- 
ditions, and securing the best results in 
practice from every part of both compli- 
cated and simple plants, this book is 
respectfully dedicated by 

THE AUTHOR. 



SECTION 1 
BOILERS 

STRENGTH OF STEAM BOILERS 

A N erroneous idea concerning the 
strength of steam boilers prevails 
among some of the men who are em- 
ployed in steam plants, which is well 
illustrated by their answers to the follow- 
ing question: "Suppose that two boilers 
are made of the same material, with riv- 
eted joints of equal strength, and are 
alike in every respect, except that one is 
48 inches and the other 72 inches in di- 
ameter. On which of them will it be 
safe to carry the higher pressure?" 

Many of the men to whom this ques- 
tion is referred reply that the larger one 
will safely carry the higher pressure, 
and the reason for this is found in the 
fact that to those who have not intelli- 
gently studied the subject the larger a 
boiler is, the greater its strength appears 
to be, which is a mistake as the following 
lines fully explain. 

While pressure acts on every inch of 
the circumference of a boiler shell, it 
operates exactly as if it pressed upon a 
space equal to the diameter, therefore if 
the shell is 72 inches in diameter and the 
pressure is 100 pounds per square inch, 



2 STEAM ENGINEERING 

the force which acts to rupture the shell 
is equal to the diameter multiplieq 
by the pressure and 72X100 = 7,200 
pounds. If we draw a circle represent^ 
ing an end view of the shell, and the: 
draw a straight line dividing it into twd 
equal parts it will be plain that the strain 
is supported by the two sides of the shellJ 
therefore each must sustain a load equa] 
to 7,200 -=-2 =3,600 pounds. 

If the shell is but 48 inches in diam-i 
eter the total load is 48X100 = 4,80C[ 
pounds, or 2,400 pounds on each side! 
The pressure required to produce a loaq 
equal to that carried on each side of tha 
72 inch shell is found by dividing 3,60(1 
by one-half of the diameter of the boilei 
in question. In this case it is 3,600^-1 
24 = 150 pounds, thus plainly demonJ 
strating that with 100 pounds on the 72 
inch shell anl 150 on the 48 inch, thq 
loads are equal. 

The iron and steel plates of whicti 
boiler shells are made, vary in strength^ 
hence it becomes necessary to know ho-w 
much stress can be safely put on thd 
plates to be used. This can only be 
determined by pulling a sample of it 
apart in a testing machine and noting 
the strain in pounds that was required 
to break it. If the plate is .5 inch thick 
and we take a strip 1 inch wide, it might 
require 30,000 pounds or more to part 



it, and although it is quite possible to do 
this it is better to adopt a plan whereby 
much less strain will answer the same 
purpose. 

If a piece of this plate 1 inch wide is 
forged down until it is .3 inch thick, its 
area will be 1X.3 = .3 square inch, 
and if pulled apart the strain required 
to do it may be 16,611 pounds. Divid- 
ing the strain required by the area of the 
test piece where it fractured, gives the 
tensile strength of the plate in pounds 
per square inch of section. In this case 
it is 16,611 -^ .3 = 55,370 pounds. 

When calculating the strength of a 
boiler, it is necessary to take into ac- 
count the thickness of the shell, as its 
strength will vary directly with its thick- 
ness. For this calculation it is cus- 
tomary to take a strip of the shell 1 inch 
wide following the entire circumference, 
determine its strength and assume that 
all other parts are equally strong. 

In order to be on the safe side, it is 
necessary to take this strip from the 
most unfavorable location, as the 
strength of any structure is determined 
by the strain that can be safely carried 
by the weakest part. 

In some parts of the shell holes are cut 
for steam nozzles and other purposes, 
therefore if the strip found in or near 
the middle of such a place should be 



4 STEAM ENGINEERING 

taken without qualification it would 
possess no strength at all, as it is cut in 
two, but in all well made boilers these 
places are strengthened by reinforcing 
rings, or by special flanging. 

The above explanation makes it plain 
that if a strip 1 inch wide is cut from a 
sheet .5 inch thick, with a tensile 
strength of 55,370 pounds, its ultimate 
strength is 55,370 X. 5 = 27,685 pounds. 

The riveted joint in the shell of a 
steam boiler can never be as strong as 
the solid plate, and as calculations for 
determining the comparative strength 
of joints require full explanation, it is 
here assumed that the design of joint 
gives .75 of the strength of the solid 
plate. 

Taking this into account we find that 
the strip 1 inch wide has strength equal 
to 55,370 X. 5 X. 75 =20,763 pounds. 

Again referring to the circle with a 
straight line drawn through the center 
of it, showing that the above calculation 
gives the strength of one side only, it is 
plain that we must calculate on one- half 
of the diameter. 

For a 72 inch boiler this is 36 inches, 
and 20,763^36 = 576 pounds, which is 
the pressure per square inch required 



to rupture this shell. As a large margin 
must be assumed for safety, a factor of 
5 will be allowed, or in other words one- 
fifth of the bursting pressure will be 
taken as the safe working pressure, and 
576-^5 = 115, showing that it is safe 
to carry 115 pounds on such a boiler. 
This explanation will make clear the 
following rule for determining the safe 
working pressure of steam boilers. 

Multiply the tensile strength of the 
plate in pounds per square inch, by its 
thickness in decimals of an inch, and by 
the comparative strength of the riveted 
joint. Divide by one-half of the diam- 
ater in inches, and the quotient is the 
bursting pressure. Divide this by the 
factor of safety (for which 5 is recom- 
mended), and the final quotient is the 
safe working pressure. 

This rule is fully explained in order 
that the reader may clearly understand 
the principles involved, and if results 
secured by it differ from other cases, the 
cause of difference may be plainly seen. 

The following table gives the safe 
working pressures of boilers from 42 to 
84 inches in diameter, with plates of 
suitable thickness, the tensile strength 
of which is stated. The table is^based 
on the foregoing rule. 



STEAM ENGINEERING 

DIAMETER OF SHELL 42 INCHES 
FACTOR OF SAFETY 5 



Tensile strength 


Tensile strength 


Tensile strength 


50.000 pounds 


55,000 pounds 


60,000 pounds 


1. 












k 
.•St 


tic 


2i 


H^ 


C/3 O 


w a 


Ho 


C/3 O 


CO a 


Ho 


W o 


w a 


25 


.60 


71 


.25 


60 78 


.25 


.60 


85 


8125 


.60 


89 


.3125 


.60 98 


.3125 


.60 


107 


375 


.70 


125 


.375 


70 137 


.375 


.70 


150 


4375 


.75 


156 


.4375 


.75 171 


.4375 


.75 


187 


5 


.75 


178 


.5 


.75 


196 


.5 


.75 


214 



DIAMETER OF SHELL 48 INCHES 
FACTOR OF SAFETY 5 



25 


.60 


61 


.25 


.60 


69 


.25 


60 


3125 


.60 


78 


.3125 


.60 


86 


.3125 


.60 


375 


.70 


109 


.375 


70 


120 


.375 


70 


4375 


.75 


136 


.4375 


.75 


150 


.4375 


.75 


5 


.75 


156 


.5 


.75 


172 


.5 


.75 



DIAMETER OF SHELL 54 INCHES 
FACTOR OF SAFETY 5 



25 


.60 56 


,25 


.60 


61 


.25 


60 


3125 


.60 69 


.3125 


.60 


76 


.3125 


.60 


375 


.70 97 


.375 


.70 


107 


.375 


.70 


4375 


.75 121 


.4375 


.75 


134 


.4375 


.75 


5 


.75 139 


.5 


.75 


153 


.5 


.75 



DIAMETER OF SHELL 60 INCHES 
FACTOR OF SAFETY 5 



.3125 


.65 68 


.3125 


.65 


74 


.3125 


.65 


81 


.375 


.70 87 


.375 


.70 


96 


.375 


.70 


105 


.4375 


.75 109 


.4375 


.75 


120 


.4375 


.75 


130 


,5 


.75 125 


5 


.75 


137 


.5 


.75 


150 


.5625 


.80 150 


.5625 


.80 


165 


.5625 


.80 


180 



DIAMETER OF SHELL 66 INCHES 
FACTOR OF SAFETY 5 



Tensile strength Tensile strength 


Tensile strength 


50,000 pounds 


55,000 pounds 


60,000 pounds 


.5 a 


1 


^1 


2,^ 


tt 








3 


W o 


CO a 


H o 


CO O 


CO a 


Ho 


CO o 


CO a 


.3125 


.65 


61 


.3125 


.65 


68 


.3125 


.65 


74 


.375 


.70 


79 


.375 


.70 


87 


.375 


.70 


95 


.4375 


.75 


100 


.4375 


.75 


109 


.4375 


.75 


119 


.6 


.75 


113 


.5 


.75 


125 


.5 


.75 


136 


.5625 


.SO 


136 


.5625 


.80 


150 


.5625 


.80 


164 



DIAMETER OF SHELL 72 INCHES 
FACTOR OF SAFETY 5 



375 


.70 73 


.375 


.70 


80 


.375 


.70 


4375 


.70 85 


.4375 


.70 


93 


.4375 


.70 


5 


.75 104 


.5 


.75 


115 


.5 


.75 


5625 


.75 117 


.5625 


.75 


129 


.5625 


.75 


625 


80 140 


.625 


.80 


153 


.625 


.80 



87 
102 
125 
141 
166 



DIAMETER OF SHELL 78 INCHES. 
FACTOR OF SAFETY 5 



375 


.70 


67 


.375 


70 


74 


.375 


.70 


4375 


.70 


78 


.4375 


.70 


86 


.4375 


.70 


5 


.75 


96 


.5 


.75 


105 


.5 


.75 


5625 


.75 


108 


.5625 


,75 


119 


.5625 


.75 


625 


.80 


128 


.625 


.80 


141 


.625 


.80 



81 
94 
115 
130 
154 



DIAMETER OF SHELL 84 INCHES 



.375 
.4375 



.5525 
.625 



FACTOR OF SAFETY 5 



.70 


62 


.375 


.70 


68 


.375 


.70 


.75 


78 


.4375 


75 


86 


.4375 


.75 


,75 


89 


5 


.75 


98 


.5 


.75 


.80 


107 


5625 


.80 


118 


.5625 


.80 


.85 


126 


625 


.85 


139 


.625 


.85 



75 
94 

107 
128 
152 



8 STEAM ENGINEERING 

As the thickness of boiler plates is frequently 
stated in sixteenths of an inch, the following table 
will enable the reader to convert it into decimal 
fractions without calculation. 



146=. 0625 
5^=. 125 

3^6=. 1875 
K=.25 



%e=.3125 »A«=.5625 
H=.375 %=.625 

%e=. 43751 H/46=. 6875 
^=.5 %=.75 



iS/ia=.8l26 

^=.875 
i%o=.9375 
1= 1000 



RULES FOR USING DECIMAL FRACTIONS 
ADDITION 

Write down the numbers to be added 
SO that each decimal point (after the 
first) , will be directly tinder the preced- 
ing, and proceed to add as with whole 
numbers. 

Example. Proof. 

17.87 .096 

.096 41.69 

3.7 17.87 

41.69 3.7 

63.356 63.356 

If the same sum is obtained when the 
numbers to be added are located in dif- 
ferent positions, it is evidence that the 
result is correct. 

SUBTRACTION 

Set down the minuend, or larger 
number, and locate the subtrahend, or 
smaller number so that one decimal 
point will be directly under the other, 
and proceed as with whole numbers. 



Example. 

107.8672 

93.7526 



14.1146 



Proof. 

93.7526 

14.1146 

107.8672 



MULTIPLICATION 

Proceed as with whole numbers and 
point off as many decimals as there are 
in the multiplicand and multiplier, or 
the first and second numbers added 
together. 

Example. Proof. 

.375 .375).23250(.62 

.62 2250 



750 


750 


2250 


750 


.23250 





DIVISION 





Divide the dividend by the divisor as 
with whole numbers, then point off as 
many decimals in the quotient or answer 
as those in the dividend exceed those in 
the divisor. 

Example. Proof. 

.235).43750(1.86 .235 

235 1.86 



2025 


1410 


1880 


1880 


1450 


235 


1410 


40 


40 


.43750 



A 



10 STEAM ENGINEERING 

In all cases where there are not enough 
figures in the answer to provide the 
required number of decimal places, 
ciphers must be added at the left hand 
until enough are secured. 

Every cipher added at the left hand 
reduces the value of the fraction, but 
when added at the right hand for con- 
venience, they do not affect the result. 

THE UNITED STATES GOVERNMENT RULE 
FOR SAFE PRESSURES 

Many engineers prefer to use this rule 
especially where they desire to carry as 
much pressure on their boilers as they 
can find authority for, and this idea is 
correct provided the rule is used in an 
intelligent manner. This rule is as 
follows : 

Multiply one-sixth of the tensile 
strength by the thickness of plate and 
divide by one-half of the diameter, for 
single riveted seams. For double riv- 
eted seams add 20 per cent. 

A very common error when consider- 
ing this rule is to conclude that a factor 
of safety of 6 is used, but this is not the 
case, because one-sixth of the strength 
of the whole plate before it is punched, 
or drilled, is taken, and as the riveted 
seams are not as strong as the full plate 
the difference between the two operate 



11 



to reduce the factor of safety. For 
illustration of this point, see the follow- 
ing example: 

Selecting from the foregoing table a 
boiler 72 inches in diameter made of 
plates .5 inch thick whose tensile 
strength is 60,000 pounds, with a double 
riveted joint that has 75 per cent, of the 
strength of the solid plate, the safe pres- 
sure is 125 pounds. As the factor of 
safety is 5, the bursting pressure is 125 
X5 = 625 pounds. 

Applying the U. S. Government rule 
to the above case gives the following 
result : 

(60,000 ^6) X.5 4- (72 ^2) =139, 
and adding 20 per cent, brings the result 
up to 166 pounds for a safe pressure. 
Dividing the bursting pressure by the 
latter, shows that the real factor of 
safety is 625 -f- 166 =3.76 although the 
apparent factor is 6. 

From the foregoing it will be plain 
that the U. S. Government rule does not 
take into consideration the varying 
strength of riveted joints of different 
designs, except that in all cases a double 
riveted joint is 20 per cent, stronger than 
the single riveted design. It is quite 
possible to design a single riveted joint 
that is stronger than another having 
two rows of rivets, but the U. S. Gov- 
ernment rule taken as a whole is not so 



12 



STEAM ENGINEERING 



short and simple as it appears, for it 
includes directions for making riveted 
joints that are strong and durable, there- 
fore the only mistake made along this 
line is when careless engineers or igno- 
rant steam users apply the rule to cases 
where infetior joints have been made by 
mcompetent boiler makers, but it was 
never intended for such applications. 

The following table has been prepared 
for the benefit of those who prefer this 
rule, but it should be remembered that 
every single riveted joint to which the 
table refers must contain 60 per cent, of 
the strength=of-the solid plate, and every 
double riveted joint must have 75 per 
cent, of it. Furthermore the factor of 
safety averages about 3.5. For these 
conditions the following safe pressures 
are allowed. 



DIAMETER OF SHELL 42 INCHES 





Tensile 


Tensile 


Tensile 




strength 


strength 


strength 




50,000 lbs. 


55,000 lbs. 


60,000 lbs. 


i^ 


'O 


«T3 


^.'^ 


S'O 


a,V. 


Ji-v 


:a5 


If 


O > 


3 > 


S > 


.s > 




Ho 


co-n 


«-C 


wn 


Q-C 


w-c 


P'C 


.25 


99 


119 


109 


131 


119 


143 


.3125 


124 


148 


136 


163 


148 


178 


.375 


148 


178 


163 


196 


178 


214 


.4375 


173 


'^.07 


191 


229 


208 


249 


.5 


198 


237 


218 


261 


238 


285 



13 



DIAMETER OF SHELL 48 INCHES 





Tensile 


Tensile 


Tensile 




strength 


strength 


strength 




■ 50,000 lbs. 


55,000 lbs. 


60,000 lbs. 


k 


^« 


<UT3 


-a 




•s" 








:3 <u 




3 OJ 






2J; 


.s > 


o > 


c > 


o > 


c > 


O > 


Ho 


m-c 


Q-n 


co"c; 


Q-n 


w"n 


Q-n 


.25 


87 . 104 


95 


114 


104 


125 


.3125 


108 130 


119 


143 


130 


156 


.375 


130 156 


143 


171 


156 


187 


.4375 


152 182 


1H7 


200 


182 


218 


.5 


173 


207 


191 


229 


208 


249 



DIAMETER OF SHELL 54 INCHES 



25 


77 


92 


85 


102 


92 


3125 


96 


115 


106 


127 


115 


375 


116 


139 


127 


152 


139 


4375 


135 


162 


148 


177 ■ 


162 


5 


154 


184 


169 


202 


185 



111 

138 
167 
194 
222 



DIAMETER OF SHELL 60 INCHES 



3125 


87 


104 


95 


114 


104 


375 


104 


125 


114 


137 


125 


4375 


120 


145 


133 


159 


146 


5 


139 


166 


153 


183 


166 


5625 


156 


187 


171 


205 


187 



125 

150 
175 
199 
224 



DIAMETER OF SHELL 66 INCHES 



3125 


79 


95 


87 


104 


95 


375 


95 


114 


104 


125 


113 


4375 


110 


132 


121 


145 


132 


5 


126 


151 


139 


166 


151 


5625 


142 


170 


156 


187 


170 



114 
136 
158 
181 
204 



DIAMETER OF SHELL 72 INCHES 



375 


87 


104 


95 


114 


104 


4375 


101 


121 


111 


133 


121 


5 


115 


138 


127 


152 


139 


5625 


130 


156 


143 


171 


156 


625 


144 


173 


159 


190 


173 



125 

145 
166 
187 
207 



DIAMETER OF SHELL 78 INCHES 



.375 


80 


96 


88 


106 


96 


115 


.4375 


93 


111 


103 


123 


112 


134 


.5 


107 


128 


117 


140 


128 


153 


.5625 


120 


144 


132 


158 


144 


172 


.625 


133 


159 


147 


176 


160 


192 



14 STEAM ENGINEERING 

DIAMETER OF SHELL 84 INCHES 





Tensile 


Tensile 


Tensile 




strength 


strength 


strength 




50,000 lbs. 


55.000 lbs. 


60,000 lbs. 


Eg 


-o 


4)T3 


V3 


■(u-d 


T3 


(UT) 




"Sj^ 


3 (U 


1^ 


3 q3 




'i?, 


x^ 




O > 


•S > 


o > 


s > 


> 


H^ 


W-S 


q;c 


WC 


Q-C 


w'C 


Q-C 


.375 


74 


89 


82 


98 


89 


107 


.4375 


86 


103 


95 


114 


104 


125 


.5 


99 


119 


109 


130 


119 


142 


.5625 


111 


133 


122 


146 


134 


160 


.625 


124 


148 


136 


163 


149 


178 



THE STRENGTH OF RIVETED JOINTS 

One of the most important problems- 
presented in connection with steam 
boilers is to calculate the strength of | 
riveted joints of various designs. As a 
general rule (or what may more prop- 
erly be called an estimate) , the strength 
of single riveted joints have 56 per cent. -: 
and double riveted joints 70 per cent, of 
the strength of the solid plate. 

However, it is not a good plan to as- 
sume this proportion of strength, for 
while some joints may greatly exceed it^ 
others will prove to be much weaker, 
hence the only safe plan is to calculate 
the strength of each joint separately 
when the safe working pressure of a 
boiler is to be determined. 

While studying this subject the prin- 
ciples involved should first receive at- 






BOILERS 15 

tention, for if they are well understood 
it will assist the engineer not only in his 
efforts to learn the rules, but also in 
remembering them afterward. 

The comparative strength of the plate 
before and after it is punched or drilled 
for the rivets, is usually determined 
first. If we take a strip of boiler plate 
3 inches wide as shown at 2 in Fig. 1, 




it represents 100 per cent, of strength, 
because none of it has been cut away 
for rivet holes. Proceeding to drill a 
%-inch hole in it, as shown at 3, it is plain 
that one-quarter of the metal has been 
cut away and it does not need any cal- 
culation to demonstrate it, but in order 
to fully illustrate the point the following 
may be considered. 

Having removed .75 inch of the metal 
there is 3 — .75 = 2.25 inches left, and 
dividing this by the original width shows 
the percentage remaining, or 3 — .75-^- 
3=.75 of the strength of the solid plate 
hence the rule. 



16 STEAM ENGINEERING 

From the pitch of the rivets subtract 
the diameter of one rivet, and divide the 
remainder by the pitch. The quotient 
is the percentage of the strength of the 
plate. 

As a formula it appears as follows: 

P= Pitch of rivets. 

D = Diameter of rivets. 

S = Percentage of strength. 

As the pitch of the rivets is used the 

strip appears with one-half of a rivet 

hole in each edge, but the result is not 

changed by it. See Fig. 2. The thick- 



ness of the plate is not considered in this 
connection as it is not necessary, and the 
use of it would make the calculation 
more complicated without correspond- 
ing benefit. 

The next point to be considered is the 
strength of [the rivets, and for conve- 



BOILERS . 17 

nience of illustration it is assumed that 
the shearing strength of the rivets and 
the tensile strength of the plate are equal. 
While this is true in some cases it is not 
in others, therefore it cannot be laid 
down as a general rule. 

The area of a circle % inch in diam- 
eter is .44 square inch, and assuming 
that there are two rows of rivets, the 
area of two as illustrated in Fig. 3 is 




FiG_» 



.44X2 = .88 square inch. As this 
must be compared with the full area of 
the plate, its thickness must now be 
taken into consideration and for this 
purpose it is assumed to be .375 inch. 
Multiplying this by the pitch of rivets 
shows that the full area is .375X3 = 
1.12 square inches. 

By dividing the area of rivets by the 
area of plate the strength of the rivets 
is found to be .88^-1.12 = .78 of the 
solid plate, hence the rule. 



18 STEAM ENGINEERING 

Divide the area of one rivet multiplied 
by the number of rows, by the thickness 
of the plate multiplied by the pitch of 
the rivets. The quotient is the percen- 
tage of strength of the solid plate. As a 
formula it appears as follows: 
AXN_, 

T X P 

A = Area of one rivet. 
N = Number of rows, 
T = Thickness of plate. 
P = Pitch of rivets. 
S = Percentage of strength. 
In the above example the strength 
would be taken as .75, because the lower 
result must always be adopted in order to 
be on the safe side. 

This example seems to show that it 
is safe to call the strength of all double 
riveted joints .70 or more, but the 
defect in it is due to the fact that it 
only illustrates plates of a low tensile 
strength, and these are not fit for the 
construction of modern boilers. 

The Hartford Steam Boiler Inspec- 
tion and Insurance Co. have adopted 
38,000 pounds per square ijich of sec- 
tional area as the shearing strength of 
rivets in single shear, and while it does 
not seem right to use an arbitrary rating 
for all joints, as rivets vary in strength, 
still if they are put into a joint in a work- 



BOILERS 19 

manlike manner they will always be as 
strong as these figures indicate, and 
where they prove to be stronger, it ren- 
ders the boiler safer. 

Adopting this standard for the 
strength of rivets, and using plates with 
a tensile strength of 55,000 pounds, the 
result is much less, as follows: 

Pitch of rivets 3 inches. 

Diameter of rivets .75 inch. 

Number of rows 2. 

Thickness of plate .375 inch. 

3-. 75 

= .75 for the plates. 



.44X2X38,000 ^^ ^ 

= .54 for the rivets. 



.375X3X55,000 

If the strength of this joint is taken 
at .70 it becomes dangerous because its 
real strength is only .54. It can be im- 
proved by reducing the pitch to 2.25 
inches. 

2.25 -.75 

— — — — = .66 for the plates. 
2.25 

.44X2X38,000 
.375X2.25X55,000 

This increases the strength of joint to 
.66 when calculated on a very conserva- 
tive basis. 



= .72 for the rivets. 



20 



STEAM ENGINEERING 



For convenience in calculating the 
strength of rivets the following table 
gives the areas of circles from >^ to 1 
inch inclusive. 



DIAMETER AND AREA OF RIVETS, 


Diameter 


Decimal 




of rivet 


equivalent 




^ 


.5 


.196 


%6 


.5625 


.248 




.625 


.306 


.6875 


.371 


M^' 


.75 


.441 


13A6 


.8125 


.518 


Vi 


.875 


.601 


1%6 


.9375 


.690 


1 


1. 


.785 



In some cases it is advisable to increase 
the pitch and add another row of rivets, 
making what is known as the triple riv- 
eted joint. The foregoing rules for 
strength of plate and rivets applies to 
such a joint and shows it to be efficient 
when well designed, but like others its 
actual efficiency should be calculated 
and not assumed. The following pro- 
portions will give good results. See 
Fig. 4. 



~\^ 



o 



y-Y 



21 



Pitch of rivets 3.5 inches. 
Diameter of rivets 13/16 inch. 
Thickness of plate .375 inch. 

3.5 -.8125 ^, , , , 

— — ' = .76 for the plates. 

0.5 

.518X3X38,000 



.375X3.5X55,000 



= .81 for the rivets. 



This demonstrates that the strength 
of such a joint may be taken at .76 of 
the strength of the solid plate. The 
same proportions may be used for a 
joint in plates of 60,000 pounds tensile 
strength, but owing to the increase in 
strength of the plates, the comparative 
strength of the rivets is less. 

.518X3X38,000 ^, , , . 
- = .75 for the rivets. 



.375X3.5X60,000 



As the comparative strength of the 
plate, when punched or drilled ready for 
use, remains the same, or .76, the strength 
of the joint must be taken at ,75 which 
is practically the same as before. 

Using the same proportions of joint 
for plates of 50,000 pounds tensile 
strength raises the comparative strength 
of rivets. 

518X3X38,000 

• = 90 

.375X3.5X50,000 



22 



STEAM ENGINEERING 



Therefore this joint may be taken at 
.75 for all plates mentioned in the fore- 
going tables, so far as their tensile 
strength is concerned. 

Modem practice in steam engineering, 
calling for high boiler pressures, has 
proved that every form of lap joint is 
unreliable, owing to the tendency to 
bend the plate as shown in Fig. 5. 




in 



and as this is repeated every time that 
pressure is raised and removed, it has 
caused disastrous failure in many cases. 



23 



To overcome this objection the double 
strap butt joint is used, a very simple 
form of which is shown in Fig. 6. Such 



a joint is very efficient even when only 
a single row of rivets is used, because 
the rivets are in double shear, or in 
other words they must be sheared in 
two places before they fail. If the 
strength in single shear is 38,000 pounds 
it is raised to 70,300 pounds for double 
shear if we consider double shear equal 
1.85 single shear. For illustration of 



24 



STEAM ENGINEERING 



the Strength of this form of joint, the 
following proportions are taken. 



Pitch of rivets 
Diameter of rivets 
Number of rows 
Thickness of plate 
Tensile strength 



3 inches. 



.375 

50,000 pounds. 



3 -.875 



.70 for the plates. 



.601X1X70,300 

•— - — - — Z7rz:z = -75 for the rivets. 

.375X3X50,000 

The strength of this single riveted 
joint is thus shown to be .70 of the 
strength of the solid plates. - 

Increasing the pitch and adding an- 
other row of rivets makes a better joint, 
for which the following proportions are 
assumed : 



Pitch of rivets 
Diameter of rivets 
Number of rows 
Thickness of plate 
Tensile strength 



3.5 inches. 



.375 " 
60,000 pounds. 



3.5— .75 
3.5 



= .78 for the plates. 



.44X2X70,300 
.375X3.5X60,000' 



= .78 for the rivets. 



25 



In this case the strength of both 
plates and rivets equals .78 of the solid 
plates. Here the tensile strength is 
60,000 pounds while in the preceding it 
is 50,000, but it may be taken at any 
less value in either case without bad 
effect on the joint, as such changes only 
raise the percentage of strength of the 
rivets. 

This joint may be still further im- 



proved by extending the inner strap as 
illustrated in Fig. 7, and adding an inner 



26 



STEAM ENGINEERING 



row of rivets as shown in Fig. 8 with 
double pitch. For illustration of the 
principles involved, the following pro- 
portions are assumed for this joint: 



j: 



B 



C 






HZ 



Pitch of outer rivets 3.5 inches. 

" " inner " 7 

Diameter of rivets % " 

Thickness of plates .375 " 

Tensile strength 60,000 pounds. 

To determine the strength of this 
joint use the following rules: From the 

pitch of the inner row of rivets (or the 



27 



double pitch), subtract the area of one 

rivet and divide the' remainder by the 

double pitch. The quotient is the 
strength of the plates. 



=L 



o 



o 



o 



o 



o 



IT 



Multiply the combined area of all 
rivets in double shear by 70,300. Mul- 
tiply the combined area of all rivets in 
single shear by 38,000 and add the two 
products together. Divide the sum by 
the strength of the section of plate as 
found by multiplying the area of plate 
between centers of the rivets, by the ten- 



28 STEAM ENGINEERING 

sile strength. The quotient is the per- 
centage of strength of the rivets. 

Stated as a formula these rules appear 
as follows: 

P = Double pitch, or pitch of the 

inner row. 
D = Diameter of rivet. 
S = Percentage of strength of plate. 

(A XR X70,300) + (E X V X38,000) 
TXPXH 

A = Area of one rivet in double shear. 
R = Number of rivets in double shear. 
E = Area of rivets in single shear. 
V = Number of rivets in single shear. 
T = Thickness of plate. 
P = Pitch of rivets. 
H = Tensile strength. 
S = Strength of rivets. 
Application of these formulas to the 
above example results as follows: 

7 -.75 
= .89 for the plates. 

(.44X4X70,300)-}-(.44XlX38,000) 

.375X7X60,000 ~ 

140,448 ^ , , . 

= .89 for the nvets. 

157,500 



BOILERS 29 

While it is not necessary to make the 
straps as thick as the plates, they must 
not be less than /^ of their thickness, 
therefore for /4 inch plates they should 
be .375 X. 625 = .234 inch or say, }4 inch 
thick. 



THE U. S. GOVERNMENT RULE FOR THE 
STRENGTH OF RIVETED JOINTS, 
IRON PLATES AND IRON RIVETS 

As these rules are similar to the fore- 
going they will be stated as formulas 
only. 

P = Pitch of rivets. 
D = Diameter of rivets. 
S = Strength of plates. 

T X P 

A = Area of rivet. 
R = Number of rows. 
T = Thickness of plates. 
P = Pitch of rivets. 
S = Strength of rivets. 

As neither the shearing strength of 
rivets, nor the tensile strength of plates 
are mentioned, it shows that they are 
equal for iron plates and iron rivets. 



30 STEAM ENGINEERING 

The above formulas apply to single, 
double and triple riveted lap joints. 

STEEL PLATES AND STEEL RIVETS 

To determine the strength of plates 
at the joint proceed as for iron plates. 

The strength of rivets is determined 
by the following formula: 
AXRX.8 
TXP 

This is the same as for iron rivets ex- 
cept that the constant .8 is added, 
because the shearing strength of steel 
rivets is taken as .8 of the tensile strength 
of the plate. 

In preceding rules and formulas the 
shearing strength of rivets is taken as a 
certain fixed value and while this may 
not seem to be correct for all cases, still 
it is the lowest value determined by 
several tests, hence a riveted joint that 
is made in a workmanlike manner should 
always show the value given and more 
in many cases. 

There is a difference between the di- 
ameter of a rivet and the diameter of the 
hole before the rivet is driven into place, 
but when the joint is finished the hole 
is or ought to be, entirely filled by the 
rivet, hence the diameter of the hole is 
used in these calculations. 



» 



BOILERS 31 

JBLE BUTT-STRAP JOINTS WITH STEEL 
PLATES AND STEEL RIVETS 
I DOUBLE RIVETED 

The percentage of strength of plate 
lis found as in preceding examples. For 
strength of rivets the following formula 
ipplies: 

AXRX1.75X.8 
TXP 
A = Area of rivets which are in 

double shear. 
R = Number of rows, which is 2 in 
this case. 
1.75= A constant for rivets in double 
shear. This represents their 
strength compared with those 
in single shear. 
.8= A constant as above men- 
tioned. 
T = Thickness of plate. 
P = Pitch of rivets. 
S = Strength of rivets. 

TRIPLE RIVETED 

Strength of plates is determined by 
the above rule for other joints. The 
strength of rivets is determined by the 
preceding rule for double riveted joints, 
except that there are 2.5 rows instead 
of 2. This assumes that the inner row 
of rivets has a double pitch or else that 



32 STEAM ENGINEERING 

the inside strap is long enough for three 
rows, and the inner strap for two rows 
of rivets. 

When designing riveted joints for 
steam boilers it is necessary to avoid 
very narrow pitches as they bring the 
rivet holes too near together for good 
results. Too wide pitches are also detri- 
mental, as the plate will spring be- 
tween the rivets, and the joint will not 
remain tight whe!n calked. 

It is sometimes claimed that all cal- 
culations give,n in books are applicable] 
to horizontal fire tube boilers only, but 
this is a mistaken idea, as they must be 
used in calculating the strength of all 
vertical fire tube boilers and many of| 
the water tube type contain steam and 
water drums that are shells carrying 
internal pressure, and are fitted with 
riveted joints the same as fire tube 
boilers. 

BRACING FLAT SURFACES 

All flat surfaces in steam boilers, ex- 
cept those of small area, require braces 
to support them under ordinary working 
pressures, therefore it becomes necessary 
to know how much pressure a given sur- 
face can safely carry. These surfaces 
are strong enough to carry a light pres- 
. sure without bracing, but this • fact is 



BOILERS 33 

not generally taken into consideration 
when calculating the strength of these 
parts, because if there is an error in such 
a plan it is on the safe side, consequently 
it can do no harm. 

It is assumed that all of the pressure 
is carried by the braces, or stay bolts 
which are only short braces, therefore, 
the safe load for them must be known 
and not exceeded in practice. 

The following rule determines the safe 
working pressure for stay bolts and 
braces where the plates are not more 
than Jl6 inch thick: 

Multiply the area of the brace by 
6,000 and divide the product by the 
horizontal multiplied by the vertical 
pitch in inches. The quotient is the 
safe working pressure. 

Analysis of this rule shows that braces 
are limited to a strain of 6,000 pounds 
per square inch of sectional area, which 
affords a liberal factor of safety as the 
tensile strength of such iron ought to be 
at least 45,000 pounds. 

Where this rule is applied to threaded 
stay bolts, the area must be calculated 
from the diameter at the base of the 
thread, as that represents the real 
strength of it. For illustration of this 
rule the following example is given: 



34 STEAM ENGINEERING 

* Diameter of stay bolt .75 inch 

Area " " " .44 square inch 

Horizontal pitch 5 inches 

Vertical pitch 6 " 

.44X6,000 ^^ 

Then =88 pounds safe 

5X6 

working pressure. 

If any form of braces are located too 
far apart, the plate between them will 
bulge and finally cause leaks or perhaps 
a disastrous rupture, consequently the 
safe load for these plates must be deter- ■ 
mined and the actual working load kept 
below the safe limit. This may be de- 
termined by the following rule, for 
plates not more than Kq inch thick. 

Multiply 28,672 by the square of the 
thickness of the plate and divide the 
product by the horizontal multiplied 
by the vertical pitch. The quotient is 
the safe working pressure. 

For example, take the water leg of a 
locomotive boiler made in the following 
proportions : 

Thickness of plate .375 inch 

Horizontal pitch 6 inches 

Vertical " 6 " 

Then 



28,672 X (.375 X .375 ) 
6X6 

pressure. 



112 pounds safe 



h£^ 



BOILERS 35 

Analysis of this rule shows that as the 
pitch increases, the safe pressure de- 
creases, which is logical. If the horizon- 
tal and vertical pitches were always 
equal the rule would read "square the 
pitch," but they sometimes differ in 
practice, hence the rule covers any differ- 
ence that is desired or required. 

Where the plates are more than 
% inch thick, the same rule applies if the 
safe pressure is based on the diameter 
and pitch of stay bolts, but larger bolts 
and greater pitches are adopted in order 
to be consistent with the increased 
thickness of plate, as the following ex- 
ample illustrates: 

Diameter of stay bolt 13^ inches. 

Area " " " 1.767 square in. 

Horizontal pitch 9.5 inches. 

Vertical " 9.5 " 

Then 

1.767X6,000 

— — = 117 pounds safe pressure. 

The following table of areas of circles 
from 1 to 2 inches in diameter is given for 
convenience in making calculations that 
are based on the foregoing rule: 



36 



STEAM ENGINEERING 



AREA OF CIRCLES FROM 1 TO 2 INCHES 



Diameter 


Decimal 


Area 


1 


1.0000 


.785 


iHe 


1.0625 


.886 


IH 


1.125 


,.994 


l^ie 


1.1875 


1.107 


IH 


1.25 


1.227 


ma 


1.3125 


1.353 


IH 


1.375 


1.484 


ITie 


1.4375 


1.622 


iy2 


1.5 


1.767 


PAe 


1.5625 


1.917 


1% 


1.625 


2.073 


11^6 


1.6875 


2.236 


IH 


1.75 


2.405 


mia 


1.8125 


2.580 


IH 


1.875 


2.761 


li%a 


1.9375 


2.948 


2 


2.0000 


3.141 



Where the safe pressure is based on 
the thickness of the plate and the pitch 
of stay bolts, the following rule applies : 

Multiply 30,720 by the square of the 
thickness of the plate and divide the 
product by the horizontal multiplied 
by the vertical pitch. 

This rule contains a larger constant 
number than the former given for this 
purpose, which is the only difference 
between them. For illustration assume 
the following proportions ; 

Thickness of plate .625 inch 
Horizontal pitch 10 inches 
Vertical " 10 

Then 

30,720 X (.625 X. 625) ^„^ ^ . 

. ^^ = 120 pounds safe 

10X10 

pressure. 



i^^ 



BOILERS 37 

The stay bolts for this case should not 
be less than 1^ inches in diameter, be- 
cause the strain to be supported is 
; lOX lOX 120= 12,000 pounds and as the 
strain on stay bolts is limited to 6,000 
pounds per square inch [of sectional 
area, these must have an area of at 
least 2 square inches, and the nearest 
to this according to the table is 1^ 
inches in diameter. 

The following table contains the safe 
loads for stay bolts and braces from % 
to 2 inches in diameter. The first col- 
umn gives the diameter, the second is 
the safe load at 6,000 pounds per square 
inch sectional area, which is the U. S. 
Government rule, but inasmuch as this 
affords a large factor of safety, the 
Hartford Steam Boiler Inspection and 
Insurance Co. allow 7,500 pounds and 
this is found in the third column. 



SAFE LOADS FOR STAY BOLTS AND 
BRACES 



Diameter 


Safe load 


Safe load 


in inches 


. 6,000 pounds 


7,600 pounds 


\i 


1,176 


1,470 


% 


1,836 


2,295 




2.646 


3,307 


.^ 


3,606 


4,507 




4,710 


5,887 


IH 


5,964 


7,355 




7,362 


9,202 


1^ 


8,904 


11.130 


1 j^ 


10.602 


13,252 


1% 


il2,438 


15,547 


15^ 


114.430 


18,037 


Wi 


16,566 


20.707 


2 


18.840 


23,557 



38 STEAM ENGINEERING 

SURFACE SUPPORTED BY STAY BOL 
AND BRACES 

Every stay bolt and brace occupies I 
more or less space on the plate that is to 
be supported, the amount varying with 
the diameter, and this space is covered, 
therefore it is not subjected to pressure. 
It is customary to ignore this fact when 
making these calculations, because if 
there is an error it is on the safe side, 
hence no harm can result. 

When braces contain toggle joints 
they should be well made, so that there 
will be no lost motion to be taken up 
when pressure is put on, and let out 
when it is removed, thus causing the 
plate to bend many times, the ulti- 
mate result of which is that the metal 
is weakened until it may fail under 
an ordinary working pressure, causing 
loss of life and property. 

The following table contains the num- 
ber of square inches on a fiat surface 
that stay bolts and braces will safely 
support, the diameter ranging from 3^ 
inch to 2 inches and the pressure from 
80 to 120 pounds. The strain is limited 



BOILERS 39 

to 6,000 pounds per square inch of sec- 
tional area. 

Following this • is another table which 
is the same except that the strain is* 
limited to 7,500 pounds per square inch 
of sectional area. 

To use these tables proceed as follows: 
Suppose that a space on the water leg 
of a locomotive boiler that is 24X56 
inches, is to be made strong enough to 
carry 90 potmds pressure and stay bolts 
"J/g inch in diameter are to be used. How 
many will be required? The space con- 
tains 24X56 = 1,344 square inches. 

The first table, limiting the stress to 
6,000 pounds, shows that each % inch 
stay bolt will support 40 square inches. 
under 90 pounds pressure. Then 1,344 
■T-40 = 33 with a remainder of 24 
inches to be provided for, therefore it 
will require 34 stay bolts. 

If the strain is limited to 7,500 pounds 
^ an examination of the next table shows 
that a J/g inch stay bolt will support 50 
square inches under 90 pounds pressure. 
Then, 1,344^50 = 26 with a remain- 
der of 44 square inches, therefore 27 
stay bolts will be required. 



40 



STEAM ENGINEERING 



SQUARE INCHES OF SURFACE 
SUPPORTED BY STAY BOLTS AND BRACES 



J_^ 


Strain Umited to 6,000 pounds 


«sl 




Boiler Pressure 




Q'SS 


80 


90 


100 


110 


120 


^ 


14.7 


13.0 


11.7 


10 .,7 


9.8 




23.0 


20.4 


18.3 


16.7 


15.3 


H 


33.0 


29.4 


26.4 


24.0 


22.0 


45.0 


40.0 


36.0 


32.7 


30.0 




58.8 


52.3 


47.1 


42.8 


39.2 


\yi 


74.5 


66.2 


59.6 


54.2 


49.7 


IH 


92.0 


81.8 


73.6 


66.9 


61.3 


ij^ 


111.3 


98.9 


89.0 


80.9 


74.2 


iH 


132.5 


117.8 


106.0 


96.3 


88.3 


1% 


155.4 


138.2 


124.3 


113.0 


103.6 


IM 


180.3 


160.3 


144.3 


131.1 


120.2 


lyi 


207.0 


184.0 


165.6 


150.6 


138.0 


2 


235.5 


209 3 


188.4 


171.3 


157.0 



SQUARE INCHES OF SURFACE 
SUPPORTED BY STAY BOLTS AND BRACES 



!r! « 


Strain Limited to 7,500 pounds 


"rt-S 




Boiler Pressure 






80 


90 


100 


110 


120 


H 


18.3 


16.3 


14.7 


13.3 


12.2 


H 


28.7 


25.5 


22.9 


20.8 


19.1 


41.3 


36.7 


33.0 


30.0 


27.5 


t/^ 


56.3 


50.0 


45.0 


40.9 


37.5 




73.5 


65.4 


58.8 


53.5 


49.0 


lyi 


91.9 


81.7 


73.5 


66.8 


61.2 


1 Ji 


115.0 


102.2 


92.0 


83.6 


76.6 


IH 


139.1 


123.6 


111.3 


101.1 


92.7 


1 ^ 


165.6 


147.2 


132.5 


120.4 


110.4 


1% 


194.3 


172.7 


155.4 


141.3 


129.5 


IH 


225.4 


200.4 


180.3 


163.9 


150.2 


IH 


258.8 


230.0 


207.0 


188.2 


172.5 


2 


294.4 


261.7 


235.6 


214.1 


196.3 



r 



BOILERS 
THE ANGULARITY OF BRACES 



41 



As a general rule, to which there may 
be a few exceptions, stay bolts are set at 
right angles to the plates, and all of the 
foregoing rules and directions are based 
on this condition, but braces must fre- 
quently be located at angles that are less 
than 90 degrees, and the effect of this is 
to put a greater load on the brace, 
although the same surface is supported 
and the same pressure carried. 

Fig. 9 illustrates a stay bolt in place 




at an angle of 90 degrees to the plate, 
therefore the pressure puts a strain on 



42 



STEAM ENGINEERING 



the brace that is represented by the sur- 
face of plate multiplied by the pressure 
per square inch. 

Fig. 10 shows a brace that is located 
at an angle of 45 degrees to the boiler 
head that it supports. The load on this 




FKS../0 



brace is determined by the surface sup- 
ported in square inches, multiplied by 
the pressure per square inch, divided by 
the natural sine of the angle. 

Suppose that this brace is IM inches 
in diameter and the strain is limited to 
6,000 pounds per square inch of section- 
al area. According to the table of safe 
loads this brace will carry 7,362 pounds 
which will be secured in accordance with 
the next table if it supports 73.6 square 
inches at 100 pounds pressure. 

The natural sine of an angle of 45 



BOILERS 43 

degrees is .7071, therefore the true load 
on this brace under these conditions is 
7,362 -=-.7071. =10,411 pounds, which 
is an overload even if the limit is placed 
at 7,500 pounds per square inch of sec- 
tional area. 

To overcome this objection it becomes 
necessary to ascertain the .number of 
square inches that can be allowed with- 
out exceeding the safe limit. This is 
accomplished by multiplying the num- 
ber of square inches that could be sup- 
ported safely if the brace stood at an 
angle of 90 degrees as stated in the table, 
by the natural size of the angle. If 
this number is 73.6 and the brace stands 
at an angle of 45 degrees, the natural 
sine of which is .7071 then the surface 
that can safely be allowed to this brace 
is 73.6X.7071=52 square inches. 
While braces are not usually located at 
such an acute angle, this example illus- 
trates the effect of angularity and shows 
that it should not be neglected where 
accuracy is desired. If the angle is in- 
creased to 75 degrees, the space becomes 
71 square inches. 

The following table contains the nat- 
ural sines of angles from 10 to 90 degrees 
which fully covers all that can be re- 
quired for use in connection with boiler 
bracing, and the use of the data given in 
it has already been fully explained. 



44 



STEAM ENGINEERING 



This is an important branch of the 
subject of boiler bracing, and is earn- 
estly recommended to those who have 
not heretofore given it due attention. 



NATURAL SINES OF ANGLES 


Angle 


Sine 


Angle 


Sine 


10 


.1736 


51 


.7771 


11 


.1908 


52 


.7880 


12 


2079 


53 


.7986 


13 


.2249 


54 


.8090 


14 


.2419 


55 


,8191 


15 


.2588 


56 


.8290 


16 


.2756 


57 


.8386 


17 


.2923 


58 


.8480 


18 


.3090 


59 


.8571 


19 


3255 


60 


.8660 


20 


.3420 


61 


.8746 


21 


.3583 


62 


.8829 


22 


.3746 


63 


.8910 


23 


.3907 


64 


.8987 


24 


.4067 


65 


.9063 


25 


.4226 


66 


.9135 


26 


.4383 


67 


.9205 


27 


.4539 


68 


.9271 


28 


.4694 


69 


.9335 


29 


.4848 


70 


.9396 


30 


:5000 


71 


.9455 


31 


.5150 


72 


.9510 


32 


.5299 


73 


.9563 


33 


.5446 


74 


.9612 


34 


.5591 


75 


.9659 


35 


.5735 


76, 


.9703 


36 


.5877 


77 


.9743 


37 


.6018 


78 


.9781 


38 


.6156 


79 


.9816 


39 


.6293 


^80 


.9848 


40 


.6427 


81 


.9876 


41 


.6560 


82 


.9902 


42 


.6691 


83 


.9925 


43 


.6820 


84 


.9945 


44 


.6946 


85 


.9961 


45 


.7071 


86 


.9975 


46 


.7193 


87 


.9986 


47 


.7313 


88 


.9993 


48 


.7431 


89 


.9998 


49 


.7547 


90 


1.0000 


50 


.7660 







BOILERS 45 

BRACING THE HEADS OF TUBULAR 
BOILERS 

Every tube that is put into the head 
of a tubular boiler acts as a brace of 
more or less efficiency, and as a large 
portion of the head is removed in mak- 
ing places for these tubes, they are suf- 
ficient to hold what is left except a small 
space below them which usually requires 
two braces, and a larger space above 
them which assumes the form of a seg- 
ment of a circle. 

It is more difficult to compute the 
area of this space than to determine the 
number of square inches on the water 
leg of a locomotive boiler, or on any 
other part that can be put into the form 
of a square, as there are several points to 
be taken into consideration. 

These heads are usually about one- 
half inch thick, and taking this as a 
basis the head is self-supporting for 2 
inches above the upper row of tubes, 
and for 3 inches from the shell, conse- 
quently if we take the head of a 72-inch 
boiler in which the upper tubes are 7 
inches above the center line, the height 
of the segment to be braced is 24 inches, 
and it is part of a circle 72 -(3X2) =66 
inches in diameter, as 6 inches must 
be subtracted for the self-supporting 
portion above mentioned. The chord 



46 



STEAM ENGINEERING 



or base is 63.375 inches long as shown 
by measurement on a full sized drawing. 
This is illustrated in Fig. 11. 




Boilers 47 

rule for calculating the area of 
a segment of a circle 

Ascertain the area of a sector of the 
circle having the same arc as the seg- 
ment. Determine the area of the tri- 
angle formed by the chord of the seg- 
ment and the radii of the sector. Sub- 
tract the latter from the former and the 
remainder is the area of the segment. 



Fig. 12 represents a sector the upper 
part of which constitutes the segment. 
Applying this rule shows that the seg- 
ment in this case contains 1,125 square 
inches. It is assumed that the head is 
full 72 inches in diameter in order to 



48 STEAM ENGINEERING 

agree with calculations made to deter- 
mine the bursting and safe working • 
pressure of boiler shells as the diameters 
given are determined by internal meas- 
urements. 

Fig. 13 illustrates another rule for 
determining the area of the segment to 
be braced, as it was laid out full size and 
divided into equal spaces 1 inch wide 
by ordinates, commencing at the center 
and working, both ways. The illustra- 
tion is reduced for convenience, but the 
correct proportions are retained. 

There are 63 ordinates and their total 
length is 1,127.7 inches, therefore the 
average height of this segment is 1,127.7 
-^63 = 17.9 inches. As it is 63.375 
inches long it contains 17.9X63.375 = 
1,134 square inches. The rule on which 
this is based may be stated as follows: 

Draw a portion of a circle that is 6 
inches smaller than the diameter of the 
boiler, to form the arc. Draw a hori- 
zontal line 2 inches above the upper row 
of tubes to form the chord. Begin at 
the center and lay off ordinates 1 inch 
apart, extending from the chord to the 
arc. Determine the length of each, add 
them together, divide by the number of 
ordinates and the quotient is the aver- 
age height. Multiply this by the length 
and the product is the number of square 
inches contained in the segment. 



49 




50 STEAM ENGINEERING 

It is difficult or impossible to secure 
exactly the same result by two different 
rules, but the difference is not sufficient 
to affect the number of braces required. 

EXPLANATION OF THE FOLLOWING TABLE 

This table contains the areas of seg- 
ments to be braced on tubular boiler 
heads from 42 to 84 inches in diameter. 
The distance from the center of the 
boiler to the top of the tubes is given, 
and it will be noted that this arrange- 
ment of tubes gives a very liberal steam 
space. If more tubes are put in the 
heating surface will be increased, but 
taken as a whole the boiler will not be 
benefited thereby. The columns ii' 
this table contain the following data: 



1. 


Diameter of boiler head. 


2. 


Diameter of circle on. which the ca' 




culation is based. 


3. 


Distance from center of head to tci 




of tubes. 


4. 


Distance from center of head d 




chord or base of segment. 


5. 


Length of chord. 


6. 


Greatest height of segment. 


7. 


Average height of segment. 


8. 


Area of segment. 


^ 





51 



All measurements are stated in inches 
and areas in square inches. • 

tabLe of areas of segments 
determined by ordinates 



1 


2 


3 


4 


5 


6 


7 


8 


42 


36 


4 


6 


34 20 


12 


8.61 


294 


48 


42 


4 


6 


40.20 


15 


10.86 


436 


54 


48 


5 


7 


46.00 


17 


12.36 


568 


60 


54 


5 


7 


52.-25 


20 


14.93 


780 


66 


60 


6 


8 


57.75 


22 


16.04 


926 


72 


66 


7 


9 


63.37 


24 


17.90 


1,134 


78 


72 


8 


10 


68.88 


26 


19.20 


1,322 


84 


78 


9 


11 


74.75 


28 


20.70 


1,547 



ANOTHER RULE 



A rule that is frequently used to de- 
termine the area of a segment of a circle, 
is expressed briefly in the following 
formula : 



H 
D 



XCXD2 = A 



H = Height of segment. 
■ D = Diameter of circle. 
C = A constant taken from a table. 
A = Area of segment. 

As the height of segment and the di- 
ameter of circle are always known or 
easily determined, this constitutes a 
very simple formula, but a table of con- 
stants is required. 



52 



STEAM ENGINEERING 



The following table gives the area of 
segments previously mentioned in order 
that the reader may compare them with 
results obtained by the plan of laying 
them out by ordinates, which is here 
considere'd the standard. 



AREA OF SEGMENTS 

DETERMINED BY THE FOREGOING 

FORMULA 



Diam. 


Height 


Constant 


Diam. 


Area 


of head 


Diam. 


42 


12/36=. 333 


.2288 


1,296 


296 


48 


i%2=.357 


.2516 


1,764 


443 


54 


i%8=.354 


.2488 


2,304 


573 


60 


20^4=.370 


.2641 


2,916 


770 


66 


22/60=. 367 


.2574 


3,600 


926 


72 


2%6=.363 


.2622 


4.356 


1,142 


78 


26,^2=. 361 


. 2555 


5,184 


1,324 


84 


28^8=. 359 


.2535 


6,084 


1,642 



The following table contains all of the 
constants that are required for ordinary 
cases along this line. The first column 
contains the quotients found by dividing 
the greatest height of segment by the 
diameter of the circle. The second 
column contains the corresponding con- 
stant. When the latter is multiplied 
by the square of the diameter of the 
circle, the product is the area of the 
segment. 



BOILERS 
TABLE OF CONSTANTS 



53 



Height 


Constant 


Height 


Constant 




Diametet 




Diameter 




.150 


.0738 


.198 


.1102 


.151 


.0745 


199 


.1110 


.152 


.0753 


.200 


.1118 


.153 


.0760 


.201 


.1126 


154 


.0767 


.202 . 


.1134 


155 


.0774 


.203 


.1142 


.156 


.0781 


.204 


.1150 


157 


.0789 


205 


.1158 


158 


.0796 


.206 


.1166 


159 


.0803 


.207 


.1174 


160 


.0811 


.208 


.1182 


.161 


.0818 


.209 


.1190 


.162 


.0825 


.210 


.1199 


163' 


.0833 


.211 


.1207 


.164 


.0840 


.212 


.1215 


.165 


.0848 


.213 


.1223 


.166 


.0855 


.214 


.1231 


.167 


.0862 


.215 


.1239 


.168 


.0870 


.216 


.1248 


.169 


.0877 


.217 


.1256 


.170 


.0885 


.218 


.1264 


.171 


.0892 


.219 


.1272 


.172 


.0900 


.220 


.1281 


.173 


.0908 


.221 


.1289 


174 


.0915 


.222 


.1297 


.175 


.0923 


.223 


.1306 


.176 


.0930 


.224 


.1314 


.177 


.0938 


.225 


.1322 


.178 


.0946 


.226 


,1331 


.179 


.0953 


.227 


.1339 


.180 


.0961 


.228 


.1347 


.181 


.0969 


.229 


.1356 


.182 


.0976 


.230 


.1364 


.183 


.0984 


.231 


.1373 


.184 


0992 


.232 


.1384 


185 


.1000 


.233 


.1390 


.186 


.1007 


.234 


.1398 


.187 


.1015 


.235 


.1406 


.188 


.1023 


.236 


.1415 


.189 


.1031 


.237 


.1423 


.190 


.1039 


.238 


.1432 


.191 


.1046 


.239 


.1440 


.192 


.1054 


.240 


.1449 


.193 


.1062 


.241 


.1458 


.194 


. 1070 


.242 


.1466 


.195 


.1078 


.243 


.1475 


196 


.1086 


.244 


.1483 


.197 


.1092 


.245 


.1492 



54 



STEAM ENGINEERING 



Table of Constants — Continued 



Height 




Height 




Diameter 


Constant 


Diameter 


Constant 


.246 


.1500 


.294 


.1926 


.247 


.1509 


.295 


.1936 


.248 


.1518 


.296 


.1945 


.249 


.1526 


.297 


.1954 


.250 


.1535 


.298 


.1963 


.251 


.1544 


.299 


.1972 


.252 


.1552 


.300 


.1981 


.253 


.1561 


.301 


.1990 


.254 


.1570 


.302 


.2000 


.255 


.1578 


.303 


.2009 


.256 


.1587 


.304 


.2018 


.257 


.1596 


.305 


.2027 


.258 


.1605 


.306 


.2036 


.259 


.1613 


.307 


.2046 


.260 


,1622 


.308 


.2053 


.261 


.1631 


.309 


.2064 


.262 


.1640 


.310 


.2073 


.263 


.1649 


.311 


.2083 


.264 


.1657 


.312 


.2092 


.265 


.1666 


.313 


.2101 


.266 


.1675 


.314 


.2110 


.267 


.1684 


.315 


.2120 


.268 


.1693 


.316 


.2129 


.269 


.1702 


.317 


.2138 


.270 


.1710 


.318 


.2148 


.271 


.1719 


.319 


.2157 


.272 


.1728 


.320 


.2166 


.273 


.1737 


.321 


.2176 


.274 


.1746 


.322 


.2185 


.275 


.1755 


.323 


.2194 


.276 


.1764 


.324 


.2204 


.277 


.1773 


.325 


.2213 


.278 


.1782 


.326 


.?.?.?.?. 


.279 


.1791 


.327 


.2232 


.280 


.1800 


.328 


.2241 


.281 


.1809 


.329 


.2250 


.282 


.1818 


.330 


.2260 


.283 


.1827 


.331 


.2269 


.284 


.1836 


.332 


.2279 


.285 


.1845 


.333 


.2288 


.286 


.1854 


.334 


.2298 


.287 


.1863 


.335 


.2307 


.288 


.1872 


.336 


.2316 


.289 


.1881 


,337 


.2326 


.290 


.1890 


.338 


,2335 


.291 


.1899 


.339 


,2345 


,292 


.1906 


.340 


,2364 


.293 


.1917 


.341 


.2364 



55 



Table of Constants — Continued 



Height 




Height 






Constant 




Constant 


Diameter 




Diameter 




.342 


.2373 


.390 


.2835 


.343 


.2383 


.391 


.2845 


.344 


.2392 


.392 


.2855 


.345 


.2402 


.393 


.2865 


.346 


.2411 


.394 


.2875 


.347 


.2421 


.395 


.2884 


.348 


.2430 


.396 


.2894 


.349 


.2440 


.397 


.2904 


.350 


.2449 


.398 


.2914 


.351 


.2459 


.399 


.2923 


.352 


• .2468 


.400 


.2933 • 


.353 


.2478 


.401 


.2943 


.354 


.2488 


.402 


.2953 


.355 


2497 


.403 


.2963 


.356 


.2507 


.404 


.2972 


.357 


.2516 


.405 


.2982 


.358 


.2526 


.406 


.2992 


.359 


.2535 


.407 


.3002 


.360 


.2545 


.408 


.3012 


.361 


.2555 


.409 


.3022 


.362 


.2564 


.410 


.3031 


.363 


.2574 


.411 


.3041 


.364 


.2583 


.412 


.3051 


.365 


.2593 


.413 


.3061 


.366 


.2603 


.414 


.3071 


.367 


.2612 


415 


.3081 


.368 


.2622 


.416 


.3091 


.369 


.2632 


.417 


.3100 


.370 


.2641 


.418 


.3110 


.371 


.2651 


.419 


.3120 


.372 


.2661 


.420 


.3130 


.373 


.2670 


.421 


.3140 


.374 


.2680 


.422 


.3150 


.375 


.2690 


.423 


.3160 


.376 


.2699 


.424 


.3169 


.377 


.2709 


.425 


.3179 


378 


.2719 


.426 


.3189 


.379 


.2728 


.427 


.3199 


380 


.2738 


.428 


.3209 


.381 


.2748 


.429 


.3219 


.382 


.2758 


.430 


.3229 


.383 


.2767 


.431 


.3239 


.384 


.2775 


.432 


.3249 


.385 


.2787 


.433 


.3259 


.386 


.2796 


.434 


.3268 


.387 


.2806 


.435 


.3278 


.388 


.2816 


.436 


.3288 


.389 


.2826 


.437 


.3298 



56 STEAM ENGINEERING 

Table of Constants — Continued 



Height 




Height 




Diameter 


Constant 


Diameter 


Constant 


.438 


.3308 


.445 


.3378 


.439 


.3318 


.446 


.3388 


.440 


.3328 


.447 


.3398 


.441 


.3338 


.448 


.3407 


.442 


.3348 


.449 


.3417 


.443 


.3358 


.450 


.3427 


.444 


.3368 







To illustrate the operation of finding 
the area of a segment by the use of this 
table, suppose that on the head of an 84 
inch boiler, there is a segment 28 inches 
high to be braced. As the circle is 84 
— 6 = 78 inches in diameter, the height 
divided by the diameter is 28-4-78 = 
,3-59 and when this number is found in 
the table, the constant opposite is ,2535. 
Squaring the diameter of the circle and 
multiplying by this constant shows that 
the area of this segment is 78X78X 
.2535= 1,542 square inches. 

DIRECTIONS FOR USING THE FOREGOING 
TABLES 

For illustration, suppose that ,on the 
head of a 72-inch boiler there is a seg- 
ment 24 inches high to be braced, to 
make it safe at 110 pounds pressure, and 
it is proposed to use round braces 1^ 
inches in diameter, or their equivalent 
in some other form. They are to be 



BOILERS 57 

located at an angle of 65 degrees from 
the head. How many braces will be 
required? 

The table on page 51 shows that if the 
head of a 72-inch boiler has a segment 
24 inches high its area is 1,134 square 
inches. The table of "Square Inches 
of Surface Supported by Stay Bolts and 
Braces, Strain Limited to 6,000 Pounds," 
shows that a brace 1}4 inches in diam- 
eter will support 66.9 square inches 
under 110 pounds pressure, if the brace 
is located at an angle of 90 degrees from 
the head. In this case it is at 65 de- 
grees, therefore as the table of "Natural 
Sines of Angles" shows that the sine of 
65 is .9063 this brace will safely support 
66.9 X. 9063 = 60.6 square inches. 

As the total area to be supported is 
1,134 square inches, it will require 1,134 
^60.6 = 18.7 braces, or in practice it 
would be called 19. If a strain of 7,500 
pounds per square inch of sectional area 
is allowed, each brace will support 83.6 
square inches at 90 degrees under 110 
poimds pressure as per table. Taking 
it at an angle^ of 65 degrees it will sup- 
port 83.6 X. 9063 = 75.7 square inches 
therefore it will require 1,134-7-75.7 = 
15 braces under these conditions. These 
tables provide for the solution of any 
problem along this line that is found 
under ordinary working conditions. 



58 STEAM ENGINEERING . 

It is not a difficult matter to adapt ' 
them to other conditions not mentioned 
directly, by the following plan. Under 
head of "Square Inches Supported by 
Stay Bolts and Braces. Strain Limited 
to 7,500 pounds," boiler pressures from 
80 to 120 pounds are included, but if a 
brace will support 18.3 square inches 
under 80 pounds pressure, it will support 
18.3 X 2 = 36.3 square inches under 
40 pounds. If it will support 200.4 
square inches under 90 pounds pressure, 
it will support 200.4^-2 = 100.2 square 
inches under 180 poimds pressure, and 
many other combinations can be secured 
from this table. 

The "Table of Areas of Segments" 
can be adapted to other sizes, as follows: 
Column 1 includes a boiler head 66 inches 
in diameter in which the distance from 
center of head to top of tubes is 6 inches. 
(See column 3.) The distance to the 
chord of this segment is 8 inches, the 
'length of the chord is 57.75 inches, and 
the area is 926 square inches. 

Suppose that in another case the 
greatest height of segment is 21 inches 
or 1 inch less than given in column 6. 
As the chord in the table is 57.75 inches 
long, the area will be nearly 57.75 square 
inches less. If the segment is 1 inch 
higher or 23 inches, its area is about 57.75 
square inches more. These results are 



BOILERS 59 

practically correct for the calculations 
to which they belong, as the slight dif- 
ference due to approximate measure- 
ments will not change the number of 
braces required in a given case. 

BOILER FLUES 

SAFE PRESSURE FOR FLUES 

Flues in a shell boiler are subject to 
collapsing pressure, hence they must be 
treated in a different way from parts 
that resist bursting pressure. When 
calculating the safe internal pressure of 
the shell of a boiler it is not necessary to 
take the length of it into consideration, 
but experiments on flues show that when 
collapsing pressure is under considera- 
tion, the length becomes a factor. The 
following rule determines the safe pres- 
sure of flues: 

Multiply the square of the thickness 
of the iron by 806,300. Divide the 
product by the diameter multiplied by 
the length and by 3. The quotient is 
the safe pressure. 

For illustration take a flue made of 
iron 34 inch thick with riveted seams. 
It is 12 inches in diameter and 20 ft. long, 

.25 X. 25X806,300 

= 70 pounds safe pres- 

12X20X3 ^ ^ 

sure. 



60 STEAM ENGINEERING 

THICKNESS OF FLUES 

The required thickness of flues for 
given conditions, is determined by the 
following rule: 

Multiply the diameter in inches, by 
the length in feet, by the steam pres- 
sure and by 3. Divide the final product 
by 806,300 and extract the square root 
of the quotient. 

The above mentioned example gives 
^his result: 



V 



20X12X70X3 

=.25 inches thick. 

806,300 



DIAMETER OF FLUES 

To determine the proper diameter of 
a flue for given conditions the following 
rule applies: 

Multiply the square of the thickness 
by 806,300. Divide the product by the 
length in feet, multipled by the steam 
pressure and by 3. The quotient is the 
diameter in inches. 

Using the foregoing, example results 
as follows: 



.25 X. 25X806,300 . , 

= 12 mches diameter. 



20X70X3 



61 



LENGTH OF FLUES 



As the length of a flue is a factor in 
determining the safe pressure, it becomes 
necessary in certain cases to determine 
how long a flue can be made without 
exceeding the safe limit. This point 
can be decided by the next rule. 

Multiply the square of the thickness 
by 806,300 and divide by the diameter 
multiplied by the safe pressure and by 3. 
The quotient is the length in feet. 

Applying this to the same example 
gives the following result: 

.25 X. 25X806,300 

■ =20 feet long. 

12X70X3 

In all the foregoing calculations con- 
cerning boiler flues, a factor of safety of 
3 is used, which seems to have proved 
sufficient for all practical use in the judg- 
ment of Inspectors in the.U. S. Govern- 
ment service, but if a greater factor is 
desired it may be substituted, giving 
results in accordance with the change. 

The length of a flue, so far as its safe 
pressure is concerned is the distance be- 
tween supports. If it is a plain flue this 
means the extreme length, but if it is 
made in sections with flanged ends, each 



62 STEAM ENGINEERING 

of these sections is called the length, 
and if re-inforcing rings are used, the 
distance between them is taken as the 
length in the above rules for safe pres- 
sure, etc. 

Furthermore, throughout this work 
a boiler flue is taken as being 43^^ inches 
or more in diameter while tubes are 4 
inches or less. In accordance with this 
assumption, the foregoing rules are not 
recommended for anything less than 5 
inches, and they probably give very 
conservative results in practice. 

There does not seem to be any really 
satisfactory data of a simple nature 
concerning the strength of boiler tubes, 
but fortunately they are strong enough 
to safely withstand much more pressure 
than is given for boiler shells of various 
diameters, consequently they are safer 
than other parts. 

The same reasoning applies to tubes 
for water tube boilers which are sub- 
jected to bursting pressure, hence they 
are properly termed "safety boilers" 
provided they do not include a large 
steam drum. 

As a general rule a water tube does 
comparatively little damage when .it 
bursts, but exceptions to this rule are 
not unknown. 

The next table contains the standard 
dimensions and weight per linear foot 



BOILERS 



'63 



(Of boiler flues from 4*^ to 21 inclies' in 
diameter. 

When referring to the size of standard 
boiler flues and tubes it is well to remem- 
ber that when the diameter is men- 
tioned it is the outside diameter, conse- 
quently when ordering flue cleaners this 
fact must be taken into consideration. 
If the internal diameter is meant in any 
special case it should be plainly stated 
in order to avoid misunderstandings. 

Following this is a table of dimensions; 
of tubes fi:om 1 to 4 inches in diameter^ 
and this refers to the external diameter 
as above mentioned. 

SIZES OF STANDARD BOILER FLUES 
FROM 4H TO 21 INCHES 



Diameter 


Thick- 


Circumference 


Weight 


Ext. 


Int. 


ness 


Ext. 


Int. 


4H 


4.232 


.134 


14.137 


13.295 


6,17 


5 


4.704 


.148 


15.708 


14.778 


7.58 


6 


5.670 


.165 


18.850 


17.813 


10.16 


7 


6.670 


.165 


21.991 


20.954 


11.90 


8 


7,670 


.165 


25.133 


.24.006 


13.65 


9 


8.640 


.180 


28.274 


27 . 143 


16,76 


10 


9.594 


.203 


31.416 


30.141 


21.00 


11 


10.560 


.220 


34.558 


33.175 


25.00 


12 


11.542 


.229 


37.699 


36.260 


28.50 


13 


12.524 


.238 


40.841 


39.345 


32.06 


14 


13.504 


.248 


43.982 


42.424 


36.00 


15 


14.482 


.259 


47.124 


45.497 


40.60 


{I 


15.468 


.271 


50.266 


48.563 


45.20 


16.432 


.284 


53.407 


51.623 


49.90 


18 


17.416 


.292 


56.549 


54.714 


54.82 


19 


18.400 


.300 


59.690 


57.805 


59.48 


20 


19.360 


.320 


62.832 


60.821 


66 77 


21 


20.320 


.340 


65.974 


63.837 


73.40 



64 



STEAM ENGINEERING 



SIZES OF STANDARD BOILER TUBES 
FROM 1 TO 4 INCHES 



Diameter 




Circumference 












Weight 








Ext. 


Int. 




Ext. 


Int. 




1 


.810 


.095 


3.142 


2.545 


.90 


IH 


1.060 


.095 


3.927 


3.330 


1.15 


1/^ 


1.310 


.095 


4.712 


4.115 


1.40 


1% 


1.360 


.095 


5.498 


4.901 


1.65 


2 


1.810 


.095 


6.283 


5.686 


1.91 


2K 


2.060 


.095 


7.069 


6.472 


2.16 


2y2 


2.282 


.109 


7.854 


7.169 


2.75 


2% 


2.532 


.109 


8.639 


7.995 


3.04 


3 


2.782 


.109 


9.425 


8.740 


3.33 


3M 


3.010 


.120 


10.210 


9.456 


3.96 


3^ 


3.260 


.120 


10.996 


10.242 


4.28 


3M 


3.510 


.120 


11.781 


11.027 


4.60 


4 


3.732 


.134 


12.566 


11.724 


5.47 



EXTRA STRONG BOILER TUBES 

There are many cases where extra 
strong tubes are required for special 
purposes and they are made to meet this 
demand. As a thick tube will hold 
more than a thin one in the capacity of 
a brace ^ they are sometimes used near 
the center of large boiler heads on this 
account. They should be used in some 
of the horizontal water tube boilers now 
in the market, in order to make them 
more reliable, as more material is needed 
in a water tube than in a fire tube for 
the same pressure, as the former must 
support the weight of water from which 
the latter is free. However, the water 
tube is supported between the ends, 



BOILERS 65 

while a fire tube cannot be so strength- 
ened. The next table gives dimensions 
of extra strong boiler tubes. 

SIZE OF EXTRA STRONG BOILER TUBES 



-w 


u 


u 


S 


*Thick- 


Inside 


"5 <u 


^l 


<u-£ 


t 


ness 


diam. 


a a 

O rt 




TS g 


^ 


double 


double 


"m S 


!3 


extra 


extra 


^--B 


O'-S 


£'3 


H 


strong 


strong 


Vi 


.84 


.542 


.149 


.298 


.244 


% 


1.05 


.736 


.157 


.314 


.422 


1 


1.315 


.951 


.182 


.364 


.587 




1.66 


1.272 


.194 


.388 


.884 


1^ 


1.9 


1.494 


.203 


.406 


1.088 


2 


2.375 


1.933 


.221 


.442 


1.491 


2H 


> 2.875 


2.315 


.280 


.560 


1.755 


3 


3.5 


2.892 


.304 


.608 


2.284 


3H 


4.0 


3.358 


.321 


.642 


2.716 


4 


4.5 


3.818 


341 


.682 


3.136 



HEATING SURFACE OF STEAM ^OILERS 

The heating surface of a steam boiler 
consists of the parts that are in contact 
with the fire or the hot gases produced 
by it, on one side, and are covered by- 
water on the other. It may be possible 
to show an exception to this rule in the 
case of a vertical fire tube, boiler in 
which the tubes extend through the 
steam space to the upper head. For a 
com.pa.ratively short distance they are 
not covered by water but they tend to 
superheat the steam, or at least to evap- 
orate particles of water that are taken 



66 STEAM ENGINEERING 

Up by the steam as it ascends from the 
water surface, hence the surface above 
the water level in such cases is of more 
or less value. 

The capacity of a boiler so far as the 
evaporation pf water is concerned, de- 
pends on the amount of heating surface 
it contains and the rapidity with which 
the water circulates through it. The 
number of square feet of heating surface 
is easily determined, but the efficiency 
of the circulation can only be determined 
by actual trial of various kinds of boilers. 

After it is once ascertained for a cer- 
tain design, all others of the same kind 
may be depended upon to show the same 
results, provided the conditions are 
alike. 

The amount of heating surface offers 
a suggestion concerning the power that 
a boiler can develop, but it should not 
be taken as definite information con- 
cerning what it can do or of what it 
actually is doing in practice. 

A boiler may have heating surface 
enough to develop 100 horse power ac- 
cording to standard rating for given 
conditions, but the actual conditions 
may be more favorable for evaporating 
water, hence it can be made to develop 
150 horse power, although it would 
probably result in low efficiency so far 
as the amotmt of water evaporated per 



BOILERS 67 

pound of coal is concerned, and repair 
bills may be large. 

On the other hand a boiler may have 
enough heating surface 'to warrant its 
rating at 100 horse power, yet owing to 
light demand for steam it may not devel- 
op more than 50 horse power, hence this 
must be taken into consideration when 
noting the amount of coal burned in 
different places. 

Some heating surface is much more 
efficient than others, but this is not con- 
sidered when the amount is stated, be- 
cause it is difficult or impossible to 
always tell just which is the most effi- 
cient, and to draw the line of separation 
between the two kinds. Where heat is 
travelling in one direction and water in 
the opposite, the best results are secured. 

When calculating the heating surface 
in the shell of a tubular boiler, or in the 
steam drum of one of the water tube 
type, multiply the circumference in feet 
by the length and divide the product by 
2: In some cases the actual surface 
with which heat comes in contact may 
exceed the result secured by the above 
rule, because the brick work may be 
drawn in to touch the shell above the 
center line, thus increasing the surface 
by a few square feet. Where accuracy 
is required it is necessary to measure 
the exact distance from the wall on one 



68 STEAM ENGINEERING 

side to a similar place on the other, fol- 
lowing the curve of the shell. This will 
give the width, and multiplying it by 
the length gives the square feet of heat- 
ing surface in the shell. 

The tubes seem to present a more 
complicated problem, but this can be 
greatly modified by modem methods. 
If we find the circumference of a tube 
in inches, then multiply it by the length 
in inches and divide the product by 144, 
the quotient will be the number of square 
feet in the tube, but it requires the use 
of many figures, and is therefore too 
long for convenience. 

The use of a table shortens the proc- 
ess, and thus proves satisfactory^ to 
the busy engineer. The following con- 
tains the required information, and 
further explanation follows it. 

Information in this table is classified 
as follows in the several columns: 

1. External diameter in inches. 

2. External area in square inches. 

3. Internal " " " " 

4. Length of tube required for one 
square foot of heating surface, inside 
measurement. 

5. Length of tube required for one 
square foot of heating surface, out- 
side measurement. 



69 



Length of tube required for one 
square foot of heating surface, based 
on the mean diameter, or the exter- 
nal diameter less the thickness. 

BOILER TUBES 



1 


2 


3 


4 


5 


6 


1 


.785 


.515 


4.479 


3.820 


4.149 


IM 


1.227 


.882 


3.604 


3.056 


3.330 




1.767 


1.348 


2.916 


2.547 


2.730 


1 5^ 


2.405 


1.911 


2.448 


2.183 


2.316 


2 


3.142 


2.573 


2.110 


1.910 


2.010 


2H 


3.976 


3.333 


1.854 


1.698 


1.770 


2H 


4.909 


4.000 


1.674 


1,528 


1.601 


2H 


5.940 


5.035 


1.508 


1.389 


1.449 


3 


7.069 


6.079 


1.373 


1.273 


1.329 


SH 


8.296 


7.116 


1.269 


1.175 


1.222 


3H 


9.621 


8.347 


1.172 


1.091 


i.l32 


3M 


11.045 


9.676 


1.088 


1.019 


1.054 


4 


12.566 


10.939 


1.024 


.955 


.990 



Suppose that a fire tube or tubular 
boiler is fitted with 90 3-inch tubes, 
16 feet long. How many square feet 
of heating surface are in the tubes ? The 
total length is 16X96 = 1,536 feet. 
As the fire can only come in contact with 
the internal surface of these tubes, the 
total length is to be divided by 1.373 
because column 4 in the foregoing table 
shows that it requires 1.373 feet in length 
to make one square foot on the tube. 
The heating surface in this case is 1,536 
-!- 1.373 =1,1 18 square feet. 

A water tube boiler contains 126 tubes 
4 inches external diameter 20 feet long. 



70 STEAM ENGINEERING 

How many square feet of heating sur- 
face do they contain? The total length 
is 20X126 = 2,520 feet. The fire comes 
in contact with the external surface of 
these tubes, hence it requires .955 feet 
in length to make one square foot. (See 
column 5 in the foregoing table.) Then 
2,520-^.955 = 2,638 square feet in these 
tubes. 

If two-thirds of the heads in the above 
mentioned tubular boiler, minus the area 
of the tubes, is considered effective heat- 
ing surface, how many square feet do 
both heads contain? 

The area of a 66-inch circle is 66X66 
X. 7854 = 3,421.19 square inches, two- 
thirds of which is 2,280.78. The exter- 
nal area of a 3-inch tube is 7.069 square 
inches. (See column 2 in the preceding 
table.) The combined area of 96 is then 
7.069X96 = 678.62. Then 2,280.78- 
678.62 = 1,602.16 square inches which is 
15.83 square feet for each, or 31.66 
for both heads. 

If it is claimed that only the internal 
area of the tubes should be subtracted, 
because the thickness of the metal ac- 
counts for the remainder, then the area 
of each tube is 6.079 square inches. (See 
column 3 in the table.) The combined 



BOILERS 71 

area is 6.079X96=583.58 square inches 
and subtracting this from two-thirds 
of the area of the head leaves 2,280.78 - 
583.58 = 1,697.20 square inches or 11.78 
square feet, or 23.56 in both heads. These 
directions apply to all boiler heads and 
any number of tubes. 

The next table gives a list of boiler 
heads from 42 to 84 inches in diameter 
the area of each in square feet, the cir- 
cumference and one-half of the circum- 
ference of the shells of equal diameter 
in feet. 

It enables the reader to readily calcu- 
late the heating surface in any boiler 
head after accounting for the tubes. 

For illustration suppose that a boiler 
head 78 inches in diameter contains 120 
three inch tubes, and the heating surface 
in it is to be determined, taking the 
whole head. The combined area of the 
tubes is 7.069X120^144 = 5.89 square 
feet. 

The table shows that this head con- 
tains 33.18 square feet and 33.18 — 
5.89 = 27.29 square feet. 

This table may also be used for deter- 
mining the heating surface in boiler 
shells. The circumference of each is 
given in feet, and it follows that this is 



72 STEAM ENGINEERING 

the number of square feet of heating 
surface for each foot in length of the 
boiler. 

For illustration suppose that a boiler 
84 inches in diameter is 20 feet long. 
How many square feet of heating 
surface are in the shell? The table 
shows that there are 21.99 for each 
foot or 21.99X20 = 439.8 in the whole 
sheU. 

If only one-half of the shell is taken 
the table shows that there are 10.99 
square feet for each foot in length, or 
219.8 square feet for one-half of the 
shell. In the case of a vertical fire tube 
boiler, the shell cannot be counted as 
heating surface because the fire does not 
touch it, but the furnace contains some 
that is very effective. Suppose that in 
a given case the circular furnace is 60 
inches in diameter. The foregoing 
table shows that the circumference of 
this furnace is 15.70 feet, or in other 
words there are 15.70 square feet for 
each foot in height, therefore, if it is 4 
feet high above the grates there are 
15.70X4 = 60.28 square feet in this 
part. The surface in the lower head 
may be determined by the rule already 
explained. 



73 



TABLE OP AREAS OF BOILER READS 

In Sqtiare Feet, the Circumference, and One -half 
the Circumference in Feet 



Diameter 


Area in 


Circum- 


One-half 
Circtim- 


in inches 


square ft. 


ference in 
feet 


ference in 
feet 


42 


9.62 


10.99 


5.49 


44 


10.55 


11.51 


5.75 


46 


11.54 


12.04 


6.02 


48 


12.56 


12.56 


6.28 


50 


13.63 


13.09 


6.54 


52 


14.74 


13.61 


6.80 


54 


15.90 


' 14.13 


7.06 


56 


17.10 


14.66 


7.33 


58 


18.34 


15.18 


7.59 


60 


19.63 


15.70 


7.85 


62 


20.96 


16.23 


8.11 


64 


22.34 


16.75 


8.37 


66 


23.75 


17.27 


8.63 


68 


25.22 


17.80 


8.90 


70 


26.72 


18.32 


9.16 


72 


28.27 


18.84 


9.42 


74 


29.86 


19.37 


9.68 


76 


31.50 


19.89 


9.94 


78 


33.18 


20.42 


10.21 


80 


34.90 


20.94 


10.47 


82 


36.67 


21.46 


10.73 


84 


38.48 


21.99 


10.99 



The heating surface in a vertical fire 
tube boiler consists of the furnace and 
the tubes. In a horizontal tubular 
boiler it includes the shell, tubes and 
heads. The locomotive type utilises 
the sides of the furnace, also the crown 
sheet, the heads and the tubes. The 
water tube kind takes all of the tubes, 
the headers and the lower half of the 
steam drum. Flue boilers include 
about two-thirds of the shell, all of the 
flues and what is left of the heads. The 



74 STEAM ENGINEERING 

cylinder only has about two-thirds of 
the shell and the same proportion of the 
heads. 

When the total heating surface in 
either or all of the above types has been 
ascertained in accordance with these 
rules, tables and suggestions, it forms a 
basis on which to make an estimate of 
what a certain kind and size of a boiler 
will do un'der fair conditions. 

The following figures show what pan 
be secured along this line. They indi- 
cate the number of square feet that will 
develop one horse power. 

HEATING SURFACE PER HORSE POWER 

Vertical 18 square feet 

Horizontal 15 " " 

Locomotive 15 " " 

Water tube 12 " " 

Flue 10 " " 

Cylinder 8 " " 

Vertical boilers are usually designed 
to include a large amount of heating sur- 
face in a comparatively small shell, but 
the vertical portion of the tubes render 
them slightly less effective than when 
in a horizontal position, as they are 
found in the common tubular and loco- 
motive boilers. 

The water tube type is more efficient 
in evaporating water per square foot of 



BOILERS 75 

heating surface, hence the number re- 
quired per horse power is less. The flue 
boiler is efficient in this respect, but 
owing to its design the heating surface 
that can be secured with a shell of given 
diameter is less than with others above 
mentioned. The cylinder boiler is cred- 
ited with developing a horse power with 
less than any of the preceding kinds, 
but the square feet of surface secured in 
a given boiler is very small, hence a 
much larger shell is required for a given 
power. 

The author is aware that a horse 
power has been developed in numerous 
cases with much less heating surface 
than these figures represent, but the 
natural tendency of steam users to load 
their machinery to the extreme limit, 
and the ambition of faithful engineers 
to meet every responsbility put upon 
them needs no encouragement here. 
On the contrary there is more need of 
conservative advice now than ever be- 
fore, and to this end it is earnestly 
recommended that boilers never be 
loaded beyond what these figures in- 
dicate. 

The worst case that has come to our 
notice in detail is where a horse power 
was actually developed for every three 
square feet of heating surface that the 
toiler contained, but the conditions 



76 STEAM ENGINEERING 

showed an utter disregard of the danger 
to human life, and indifference to the 
cost of maintenance and repairs. 

ACTUAL HORSE POWER OF BOILERS 

When men who are interested in 
steam engineering consider the subject 
of the power of a boiler, they seem to 
naturally divide themselves into three 
classes. 

First, those who claim that there is 
no such thing as the horse power of a 
boiler. 

Second, those who admit that a boiler 
does develop power, but think that the 
present way of determining it is illog- 
ical and wrong, hence wish to have a 
different standard adopted. 

Third, those who admit that the pres- 
ent practice along this Une is apparently 
not up-to-date, but cannot understand 
how any standard can be devised which 
will suit even a majority of the cases in 
common practice, and so long as this is 
true now and liable to be for a long time 
to come, it is useless to introduce a decid- 
edly disturbing element into the theory 
of steam engineering, which cannot be 
of sufficient advantage to recompense 
for the trouble and expense of changing 
set rules in thousands of volumes on this 
important subject. 



BOILERS 77 

In reply to the first it is proper to call 
attention to the fact that all standards 
of this or any other kind are simply the 
result of the deliberations of one man that 
subsequently were adopted by men in 
council, or they were accepted by bodies 
of men congregated for the purpose, 
hence have become laws on the various 
subjects to which they apply. 

It is just as logical and proper to de- 
cide on what shall be known as one 
horse power in a boiler, as it is to 
decide that 5,280 feet shall constitute 
a mile. 

Concerning the objections of the sec- 
ond class, it is only necessary to say that 
the present standard was adopted many 
years ago when the average boiler pres- 
sure was much lower than it is at the 
present time. However, even at that 
time there were many boilers that were 
operated imder much lower pressure 
than the standard calls for, while others 
carried higher pressures. 

Now the only difference between con- 
ditions at that time and at the present 
is that some boilers are carrying much 
higher pressure than ever before. On 
the other hand thousands of boilers are 
operated every day under less pressure 
than the standard calls for, but this 
standard can be appHed to every boiler 
in use, without regard to the conditions 



78 STEAM ENGINEERING I 

under which it is operated. Could any 
other standard do more? 

The third class are a conservative 
body of men who are willing to accept 
anything that is decidedly better than 
what has been, or is now in use, but they 
insist on being convinced that old stand- 
ards are defective, and that new ones 
are free from defects before they discard 
the former and adopt the latter. Sure- 
ly this is safe ground on which to stand, 
and without this element the whole 
superstructure of steam engineering 
would become unstable and imreliable 
so far as theoretical standards are con- 
cerned, and this means almost every 
branch of the subject. 

The foregoing remarks refer to what 
is known as the Centennial rating of the 
power of steam boilers, as follows: 

The evaporation of 30 pounds of water 
per hotu, from feed water at 100 degrees 
Fah. into steam at 70 pounds gauge pres- 
sure, or its equivalent under other con- 
ditions, constitutes one horse power. 

If boilers were always nm under these 
conditions, it would only be necessary 
to divide the weight of water evaporated 
per hoiu- by 30 and the quotient would 
be the power developed. In practice it 
becomes necessary to reduce the actual 
results seciu-ed, to terms of the above 
rating in order to make it agree, and this 



A 



BOILERS 79 

proves a stumbling block to the working 
engineer and the owners of boilers who 
have not given the subject due atten- . 
tion. The matter will be clear to all 
who study the following statements: 

THE OBJECT OF A BOILER TEST 

The object sought in conducting a 
boiler test is to accurately determine the 
amount of water evaporated into dry 
steam, in order to calculate the power 
developed, and to find the weight of coal 
actually burned in order to ascertain 
the efficiency of the boiler. 

The foregoing seem to be self-evident 
facts, but past experience shows that 
they are not, for numerous so-called tests 
have been reported which neither tell 
how much water was evaporated nor 
the weight of coal burned to evaporate 
it. Where these points are not settled, 
the whole proves tinreliable and unsat- 
isfactory. 

Reports of conditions that can have 
no effect on the results will not be intro- 
duced for the purpose of making the 
matter appear more complicated, as the 
object is to simplify the process and 
retain accuracy. 



80 STEAM ENGINEERING 



STARTING A TEST 



Before a test is started the water level • 
' should be brought to the point where it 
is considered advisable to carrv' it which 
will be at about two gauges usually. Tie 
a string around the gauge glass at this 
point, maintain the same water level as 
closely as possible throughout the test, 
and bring it to exactly the same place 
at the conclusion of it. As a further 
precaution the steam pressure should 
be the same at the beginning and the 
end of the test. 

If these directions are followed there 
will be no error on account of more 
or less water in the boiler than there 
should be. 

Care should be taken to know that 
none of the water pumped in is lost 
through a leaky blow off valve, or at any 
other point below the water line. 

It is very convenient to use a water 
meter to determine the amount deliv- 
ered to the boiler, and there is no good 
reason why this method should not be 
used. If a water meter does not cor- 
rectly indicate the amount of water pas- 
sing through it, and its indications are 
accepted without correction, the final 



^ 



BOILERS 81 

result will not be correct, but it is not so 
difficult to calibrate, or prove a meter as 
it appears. 

Set a barrel on a pair of platform 
scales and note its exact weight. Let 
water pass through the meter into the 
barrel until the meter indicates that say, 
4 cubic feet have passed. Put a ther- 
mometer into this water and note its 
temperature, which is assumed to be 68 
degrees. 

By referring to a table of the proper- 
ties of water the weight per cubic foot at 
68 degrees is found to be 62.33 poimds, 
therefore 4 cubic feet weighs 62.33X4 
= 249.32 pounds. Suppose that the 
water in the barrel actually weighs 
244.25 pounds. In that case the water 
actually delivered to the boiler is 244.25 
H- 249.32 = .98 of what the meter in- 
dicates, consequently when the meter 
is read at the end of the test, the result 
must be multipHed by .98 to ascertain 
the true quantity in cubic feet. 

Suppose that the water actually 
weighed 253 pounds, then the indica- 
tions of the meter would "have to be 
multiplied by 253-^249.32 = 1.015 in 
order to determine the actual quantity 
used. 



82 



STEAM ENGINEERING 



WEIGHING THE WATER 



Water for testing an ordinary boiler 
may be weighed by means of three bar- 
rels as shown in Fig. 14. Two of them 




are mounted on a platform and are fitted 
with outlet pipes 2 inches in diameter, 
both of which discharge into another 
barrel under them. Valves are pro- 
vided to shut off either one at pleasure. 

Before the lower barrel is placed in 
position and connected to the pump, 
both of the upper ones should be filled 



BOILERS 83 

with water and each of them drained 
separately into a barrel on a pair of 
scales, thus determining the exact 
weight of water that each holds. 

The third barrel is then placed in posi- 
tion and connected as shown, so that 
during the test water may be drawn 
continuously from the lower barrel as it 
is fed from each of the others alternate- 
ly. By keeping a record of the number 
of times these barrels are filled and 
emptied, the exact weight of water* used 
may be known without further calcu- 
lation. 

QUALITY OF STEAM PRODUCED 

A boiler test made without determin- 
ing the quality of the steam produced, 
is of no value, because much of the water 
pumped into the boiler may not be evap- 
orated, but pass away with the steam 
in the form of hot water. 

This water may be raised to a temper- 
ature equal to the steam with which it 
mingles, and the boiler should be cred- 
ited with this heat, but the latent heat 
of evaporation has not passed into it, 
and as this is much greater than the sen- 
sible heat, it cannot be ignored. 

A sample of the steam to be tested 
should be taken from a vertical pipe if 
possible, by means of a small perforated 



84 



STEAM ENGINEERING 



pipe screwed into it as shown in Fig. 15. 
It should always be arranged so as to 
prevent taking steam from the inner 




FKS.I5 

surface of the large pipe, as water may 
trickle down on this surface, thus not 
giving a true sample of the steam. 

If it is necessary to take the sample 
from a horizontal pipe, special care 
should be taken to avoid the water 
which alwa3^s runs along the bottom of 
such a pipe. 

Having property connected this pipe 
the steam may be tested by blowing a 



BOILERS 85 

portion of it into a certain weight of cold 
water, as illustrated in Fig. 16. A tee 
should be put on the end of this pipe to 




prevent the steam from interfering with 
correct weighing of the whole. The 
temperature of this water should be 
raised to not less than 110, nor more 
than 150 degrees Fah. The former 
ought to be secured in order to give 



_^ 



86 STEAM ENGINEERING 

sufficient rise in temperature to lessen 
the possibility of error, and it is a good 
idea not to exceed the latter, as the dan- 
ger of loss of heat increases with the 
temperature. It should be thoroughly- 
stirred to make the temperature even 
throughout the whole of it. 

It is not necessary to use any given 
weight of cold water (provided the cor- 
rect weight is known), but from 300 to 
400 pounds is suggested, because more 
reliable results are usually secured with 
a large quantity. An ordinary oil bar- 
rel will hold 320 pounds without being 
too full for convenience, and as one of 
them is usually available around a steam 
plant it can be cleaned and used for this 
purpose, therefore a uniform weight of 
320 pounds is recommended for these 
tests. 

The percentage of moisture in the 
steam tested may be determined by the 
following formula: 



^X(H-C)-(T-H) 
o 



r=» 



W = Weight of cold water used. 
S = « « steam condensed. 
H = Total heat of one pound of the 

heated water. 
C = Total heat of one pound of the 

cold water. 



BOILERS 87 

T = Total heat of one pound of the 
water at a temperature cor- 
responding to the pressure of 
steam used. 
L = Latent heat of the steam, 
Q= Quality of the steam tested, 
taking dry steam at unity or I. 
The weight of steam condensed is 
found by subtracting the weight of the 
cold water from the weight of the heated 
water. 

In all cases where there is moisture in 
the steam tested, the value of Q is less 
than one, as it represents the compara- 
tive value of the steam, as for illustra- 
tion, if the formula is applied to a given 
case and the value of Q is .95 the steam 
is .95 dry or I -Q= I -.95 = .05 moist, 
or in other words it is 95 per cent, dry 
and there is 5 per cent, of moisture in it. 
For illustration, suppose that 320 
pounds of water are put into the barrel 
at a temperature of 65 degrees Fah., the 
total heat of which is 33.01. Steam is 
blown into this water until it weighs 
340 pounds or 340-320 = 20 pounds 
more and the temperature is raised to 125 
degrees Fah., the total heat of which is 
93.17. Pressure is maintained at 95 
pounds absolute, or 80 by the gauge. The 
total heat of the water under this pres- 



66 STEAM ENGINEERING i 

r 

sure is 295.1 and the latent heat of the It 
steam is 885.6. 

An application of the formula results 
as follows: 



320 

X (93.17 - 33.1) - (295.1-93.17) 



= (16X60.07 -201.93) X. 0011.= .83 

885 .6 



Therefore .83 of the mixture coming 
from the boiler is dry steam and 1.00 — 
.83 = .17 of it is water. If it only re- 
quired 19 pounds of condensed steam to 
secure the same temperature in the bar- 
rel, then the quality of the steam would 
be .89 and if 18 pounds were sufficient it 
would be raised to .94. As 17.5 pounds 
will raise it to .98 it illustrates the neces- 
sity of taking great care to secure cor- 
rect weights for every experiment. 

In case that only 17 pounds of con- 
densed steam gives the required rise in 
temperature imdef the same conditions 
the final result is 1.02 and as this is more 
than I, further explanation is necessary. 
It denotes that the steam is superheated 
slightly. 

It now becomes necessary to proceed 
as follows, in order to determine the de- 
grees of superheat. From the quality 



BOILERS 89 

of the steam subtract 1 and multiply 
the remainder by 2.0833. In this 
case it is 1.02-1X2.0833 = .04 degree, 
showing that the steam is practically 
dry. 

Suppose that it required only 5 
pounds of water to secure the same rise 
in temperature, then there would be 6 
degrees of superheat present. The tem- 
perature of this steam as it comes from 
the boiler would be 6 degrees higher 
than the temperature of saturated or 
dry steam at the same pressure. 

THERMOMETERS 

In order to make the necessary cal- 
culations after testing the steam, a 
table of the properties of water is re- 
quired, and it follows this explanation. 
The temperatures are given from the 
freezing to the boiling point on the 
Fahrenheit, Centigrade and Reaumur 
thermometers, to enable the reader to 
easily use the kind that is preferred. It 
should be remembered, however, that zero 
on the Fahrenheit scale is 32 degrees 
below the freezing point, while the 
boiling point under atmospheric pressure 
at sea level is 212. Zero and the freezing 
point are the same on the Centigrade scale 
and the boiling point is foimd at 100, 
while the Reaumur takes the freezing 



90 STEAM ENGINEERING 

point at zero and places the boiling point: 
at 80 degrees above it. 

For temperatures higher than the 
table contains the following rules may- 
be used to convert the value of one scale 
into a corresponding value on another. 

TO CHANGE FAHRENHEIT TO CENTIGRADE 

Subtract 32, multiply the remainder 
by 5 and divide the quotient by 9. Ex- 
ample: A barrel contains 320 pounds 
of water at 110 degrees Fahrenheit. 
What is its temperature by the Centi- 
grade scale? 

110-32X5^9 = 43.3 degrees. 

TO CHANGE FAHRENHEIT TO REAUMUR 

Subtract 32, multiply the remainder 
by 4 and divide the quotient by 9. 

In the foregoing example what is the 
temperature of the water by the Reau- , 
mur scale? 

110-32X4^-9=347 degrees. 

TO CHANGE CENTIGRADE TO FAHRENHEIT 

Multiply by 9, divide the product by 
5 and add 32 to the quotient. Exam- 
ple: A barrel contains 320 pounds of 



BOILERS 91 

water at 70 degrees Centigrade. What 
is its temperature by the Fahrenheit 
scale? 

70X9 -^5+32 = 158 degrees. 

TO CHANGE CENTIGRADE TO REAUMUR 

Multiply by 4 and divide the product 
by 5. 

In the preceding example what is the 
temperature of the water by the Reau- 
mur scale? 

70X4^5 = 56 degrees. 

TO CHANGE REAUMUR TO CENTIGRADE 

Multiply by 5 and divide the product 
by 4. Example: A barrel contains 
320 pounds of water at 60 degrees Reau- 
mur. What is its temperature on the 
Centigrade scale? 

60X5-^4 = 75 degrees. 

TO CHANGE REAUMUR TO FAHRENHEIT 

Multiply by 9, divide the product by 4 
and add 32 to the quotient. In the pre- 
ceding example what is the temperature 
of the water on the Fahrenheit scale? 

60X9-^4^-32 = 167 degrees. 

These rules may be applied to any 
case with correct results. 



92 STEAM ENGINEERING 



PROPERTIES OF WATER 



Temperature 


Heat 


Weight per 


F 


c 


R 


units 


cubic foot 


32 











62.42 


33 


0.6 


0.4 


1 


62.42 


34 


1.1 


0.9 


2 


62.42 


35 


1.7 


1.3 


3 


62.42 


36 


2.2 


1.8 


4 


62.42 


37 


2.8 


2.2 


5 


62.42 


38 


3.3 


2.7 


6 


62.42 


39 


3.9 


3.1 


7 


62.42 


40 


4.4 


3.6 


8 


62.42 


41 


5.0 


4.0 


9 


62.42 


42 


5.6 


4.4 


10 


62.42 


43 


6.1 


4.9 


11 


62.42 


44 


6.7 


5.3 


12 


62.42 


45 


7.2 


5.8 


13 


62.42 


46 


7.8 


6.2 


14 


62.42 


47 


8.3 


6.7 


15 


62.42 


48 


8.9 


7.1 


16 


62.41 


49 


9.4 


7.6 


17 


62.41 


50 


10.0 


8.0 


18 


62.41 


51 


10.6 


8.4 


19 


62.41 


52 


11.1 


8.9 


20 


62.40 


53 


11.7 


9.3 


20.01 


62.40 


54 


12.2 


9.8 


22.01 


62.40 


55 


12.8 


10.2 


23.01 


62.39 


56 


13.3 


10.7 


24.01 


62.39 


57 


13.9 


11.1 


25.01 


62.39 


58 


14.4 


11.6 


26.01 


62.38 


59 


15.0 


12.0 


27.01 


62.38 


60 


15.6 


12.4 


28.01 


62.37 


61 


16.1 


12.9 


29.01 


62.37 


62 


16.7 


13.3 


30.01 


62.36 


63 


17.2 


13.8 


31.01 


62.36 


64 


17.8 


14.2 


32.01 


62.35 


65 


18.3 


14.7 


33.01 


62.34 


66 


18.9 


15.1 


34.02 


62.34 


67 


19.4 


15.6 


35.02 


62.33 


68 


20.0 


16.0 


36.02 


62.33 


69 


20.6 


16.4 


37.02 


62.32 


70 


21.1 


16.9 


38.02 


62.31 


71 


21.7 


17.3 


39.02 


62.31 


72 


22.2 


17.8 


40.02 


62.30 


73 


22.8 


18.2 


41.02 


62.29 


74 


23.3 


18.7 


42.03 


62.28 


75 


23.9 


19.1 


43.03 


62.28 


76 


24.4 


19.6 


44.03 


62.27 



93 



Properties of Water — Continued 



Temperature 


Heat 


Weight per 


F 


C 


R 


units 


cubic foot 


77 


25.0 


20.0 


45.03 


62.26 


78 


25.6 


20.4 


46.03 


62.25 


79 


26.1 


20.9 


47.03 


62.24 


80 


26.7 


21.3 


48.04 


62.23 


81 


27.2 


21.8 


49.04 


62.22 


82 


27.8 


22.2 


50.04 


62,21 


83 


28.3 


22.7 


51.04 


62.20 


84 


28.9 


23.1 


52.04 


62.19 


85 


29.4 


23.6 


53.05 


62.18 


86 


30.0 


24.0 


54.05 


62.17 


87 


30.6 


24.4. 


55.05 


62.16 


88 


31.1 


24.9 


56.05 


62.15 


89 


31.7 


25.3 


57.05 


62.14 


90 


32.2 


25.8 


58.06 


62.13 


91 


32.8 


26.2 


59.06 


62.12 


d2 


33.3 


26.7 


60.06 


62.11 


93 


33.9 


27.1, 


61.06 


62.10 


94 


34.4 


27.6 


62.06 


62.09 


95 


35.0 


28.0 


63.07 


62.08 


96 


35.6 


28.4 


64.07 


62.07 


97 


36.1 


28.9 


65.07 


62.06 


98 


36.7 


29.3 


66.07 


62.05 


99 


37.2 


29.8 


67.08 


62.03 


100 


37.8 


30.2 


68.08 


62.02 


101 


38.3 


30.7 


69.08 


62.01 


102 


38.9 


31.1 


70.09 


62.00 


103 


39.4 


31.6 


71.09 


61.99 


. 104 


40.0 


32.0 


72.09 


61.97 


105 


40.6 


32.4 


73.10 


61.96 


106 


41.1 


32.9 


74.10 


61.95 


107 


41.7 


33.3 


75.10 


61.93 


108 


42.-2 


33.8 


76.10 


61.92 


109 


42.8 


34.2 


77.11 


61.91 


110 


43.3 


34.7 


78.11 


61.89 


111 


43.9 


35.1 


79.11 


61.88 


112 


44.4 


35.6 


80.12 


61.86 


113 


45.0 


36.0 


81.12 


61.85 


114 


45.6 


36.4 


82.13 


61.83 


115 


46.1 


36.9 


83.13 


61.82 


116 


46.7 


37.3 


> 84.13 


61.80 


117 


47.2 


37.8 


85.14 


61.78 


118 


47.8 


38.2 


86.14 


61.77 


119 


48.3 


38.7 


87.15 


61.75 


120 


48.9 


39.1 


88.15 


61.74 


121 


49.4 


39.6 


89.15 


61.72 


122 


50.0 


40.0 


90.16 


61.70 



04 STEAM ENGINEERING 

Properties of Water — Continued 



Temperature 


Heat 

units 


Weight per 


F 


C 


R 


cubic foot 


123 


50.6 


40.4 


91.16 


61.68 


124 


51.1 


40.9 


92.17 


61.67 


125 


51.7 


41.3 


93.17 


61.65 


126 


52.2 


41.8 


94.17 


61.63 


127 


52.8 


42.2 


95.18 


61.61 


128 


53.3 


42.7 


96.18 


61.60 


129 


53. « 


43.1 


97.19 


61.58 


130 


64.4 


43.6 


98.19 


61.66 


131 


55.0 


44.0 


99.20 


61.54 


132 


65.6 


44.4 


100.20 


61.62 


133 


66.1 


44.9 


101.21 


61.51 


134 


56.7 


45.3 


102.21 


61.49 


135 


57.2 


45.8 


103.22 


61.47 


136 


67.8 


46.2 


104.22 


61.45 


137 


68.3 


46.7 


105.23 


61.43 


138 


58.9 


47.1 


106.23 


61.41 


139 


59.4 


47.6 


107.24 


61.39 


140 


60.0 


48.0 


108.25 


61.37 


141 


60.6 


48.4 


109.25 


61.36 


142 


61.1 


48.9 


110.26 


61.34 


143 


61.7 


49.3 


111.26 


61.32 


144 


62.2 


49.8 


112.27 


61.30 


145 


62.8 


60.2 


113.28 


61.28 


146 


63.3 


60.7 


114.28 


61.26 


147 


63.9 


61.1 


115.29 


61.24 


148 


64.4 


51.6 


116.29 


61.22 


149 


65.0 


52.0 


117.30 


61.20 


150 


65.6 


52.4 


118.31 


61.18 


151 


66.1 


62.9 


119.31 


61.16 


152 


66.7 


53.3 


120.32 


61.14 


153 


67.2 


53.8 


121.33 


61.12 


154 


67.8 


64.2 


122.33 


61.10 


155 


68.3 


64.7 


123.34 


61.08 


156 


68.9 


55.1 


124.35 


61.06 


157 


69.4 


65.6 


125.35 


61.04 


158 


70.0 


66.0 


126.36 


61.02 


159 


70.6 


56.4 


127.37 


61.00 


160 


71.1 


56.9 


128.37 


60.98 


161 


71.7 


57.3 


129.38 


60.96 


162 


72.2 


57.8 


130.39 


60.94 


163 


72.8 


58.2 


131.40 


60 92 


164 


73.3 


68.7 


132.41 


60.90 


165 


73.9 


59.1 


133.41 


60.87 


166 


74.4 


59.6 


134.42 


60.85 


167 


75.0 


60.0 


135.43 


60.83 







BOILERS 


95 




Properties of Water — Continued 


Temperature 


Heat 


Weight per 


F 


C 


R 


units 


cubic foot 


168 


75 6 


60.4 


136 44 


60.81 


169 


76 1 


60 9 


137.45 


60.79 


170 


76 7 


61 3 


138.45 


60.77 


171 


77 2 


61 8 


139 46 


60.75 


172 


77 8 


62,2 


140 47 


60.73 


173 


78.3 


62.7 


141 48 


60.70 


174 


78.9 


63.1 


142 49 


60.68 


175 


79 4 


63.6 


143.50 


60.66 


176 


80.0 


64.0 


144 51 


60.64 


177 


80.6 


64 4 


145.52 


60.62 


i 178 


81 1 


64 9 


146.52 


60.59 


179 


81 7 


65.3 


147.53 


60.57 . 


180 


82 2 


65.8 


148.54 


60.55 


181 


82.8 


66.2 


149.55 


60.53 


182 


83.3 


66.7 


150 56 


60.52 


183 


83.9 


67 1 


151.57 


60.48 


184 


84.4 


67 6 


152 58 


60.46 


185 


85 


68.0 


153 50 


60.44 


• 186 


85 6 


68 4 


154.60 


60.41 


187 


86 1 


68 9 


155 61 


60.39 


188 


86.7 


69 3 


156.62 


60.37 


189 


87 2 


69.8 


157 63 


60.34 


190 


87 8 


70 2 


158.64 


60.32 


191 


88 3 


70 7 


159 65 


60.29 


192 


88 9 


71 1 


160.67 


60.27 


193 


89 4 


71.6 


161.68 


60.25 


194 


90 


72 


162 69 


60.22 


195 


90 6 


72.4 


163 70 


60.20 


196 


91 1 


72.9 


164.71 


60.17 


197 


91 7 


73.3 


165 72 


60.15 


198 


92 2 


73.8 


166 73 


60.12 


199 


92 8 


74 2 


167. 74 


60.10 


200 


93.3 


74 7 


168.75 


60.07 


201 


93 9 


75 1 


169.77 


60.05 


202 


94 4 


75.6 


170. 78 


60 02 


203 


95.0 


76.0 


171 79 


60.00 


204 


95.6 


76 4 


172.80 


59.97 


205 


96.1 


76 9 


173.81 


59.95 


206 


96 7 


77.3 


174.83 


59.92 


207 


97 2 


77.8 


175.85 


59.89 


208 


97 8 


78.2 


176.85 


59.87 


209 


98 3 


78 7 


177.86 


59.84 


210 


98.9 


79 1 


178 87 


59.82 


211 


99.4 


79 6 


179.89 


59.79 


212 


100.0 


80 


180.90 


59.76 



96 STEAM ENGINEERING 

The foregoing tables show that as the 
temperature of water is increased it ex- 
pands, hence if a cubic foot at 40 de-' 
grees Fah. or at any higher temperature 1 
is still further heated, it will occupy 
more than a cubic foot of space, and it 
follows that if the volume is kept con- 
stant the weight must decrease. 

The difference is so small that it doc ; 
not appear in the table because the fif 
ures are given to two decimal place, 
only, until the temperature is raised to 
48 degrees or more. If it is raised from 
32 to 39 degrees it contracts slightly, 
thus occupying less space, but the differ- 
ence is very small. Maximum density 
is attained at 39.1 degrees Fahrenheit, 
3.9 Centigrade or 3.1 Reaumur. 

It is useless for engineers, under even 
the best working conditions, to attempt 
to prove these figures, on account of dif- 
ficulty of measuring out exactly a cubic 
foot, and of obtaining a sample of per- 
fectly pure water. 

SATURATED STEAM 

This name does not always seem to 
be appropriate, as it suggests to the 
average working engineer, steam that is 
saturated with water, whereas it is in- 
tended for dry steam, as otherwise it 
could not constitute a standard for com- 



BOILERS 97 

parisori, for as soon as steam becomes 
mixed with water, or when there is water 
suspended in the steam, its quaHty be- 
comes variable with the amount of water 
present, hence it could not be used with 
profit for comparison. Saturated steam 
is therefore the dividing quality between 
wet and superheated steam. 

In order to solve probleras that have 
already been presented, as well as some 
of those that follow, it is necessary to 
know something of the properties of sat- 
urated steam. These are given in the 
next table which will be explained here 
in order that the working engineer who 
has not enjoyed the advantages of a 
technical education may understand 
them, also for firemen who wish to be 
advanced to engineers, and all others 
interested in this important subject. 

Absolute pressures are used because 
all pressure must be reckoned from a 
perfect vacuum in this work, as other- 
wise there would be no standard for a 
base of operations. For all practical pur- 
poses it may be taken at 15 pounds above 
the gauge pressure, for steam gauges 
indicate the unbalanced pressure, and 
safety valves are designed to operate on 
the same principle. To ascertain the 



98 STEAM ENGINEERING 

gauge pressures, subtract 15 from those 
given in the table. 

Temperatures are stated in the Fah- 
renheit scale in which zero is 32 degrees 
below the freezing point. It is not con- 
venient or practicable to state the heat 
units in water above zero, because the 
amount of heat required to raise the 
temperatures through a given range 
below 32 on this scale is less than to 
raise it through the same range above 
32 degrees. It is therefore better to 
base all such calculations in which heat 
units are factors, on the freezing point, 
as it saves confusion and trouble. 

It is claimed that the temperature of 
water and of steam in a boiler is the 
same, and this is true after the water 
has been in the boiler under working 
conditions long enough to attain its 
maximum temperature, but it does not 
immediately flash into steam, because 
it lacks the latent heat of evaporation. 
This is a wise provision as otherwise we 
could not operate steam boilers as we 
do at present. When the heat units 
in a pound of water and the latent 
heat of evaporation are added, the sum 
is the total heat of steam at given 
pressure. 



We are sometimes told that the total 
1 heat of steam is the same for all pres- 
sures, but this is not true, hence should 
not be accepted. 

The latent heat decreases as the tem- 
perature or sensible heat increases, but 
not in the same proportion as the total 
for 15 pounds, absolute pressure, is' 
1,146.9 while for 200 pounds it is, 1,198.3 
a difference of 51.4 heat units, or more 
than 4 per cent. 

It is necessary to know the weight of 
steam under various pressures when 
calculating the weight of a given volume 
and this is stated in the table. 

PROPERTIES OF SATURATED STEAM 

P= Absolute pressure in pounds per 
square inch, or 15 pounds above 
gauge pressure. 

T = Temperature of steam under the 
given pressure. 

W = Heat units in a pound of water 
under pressure that corresponds 
to the temperature. 

L= Latent heat, or the number of heat 
imits required to convert one 
pound of water at a given tem- 
perature and pressure into steam. 

S = Total heat of steam per pound 
above 32 degrees Fah. 

C= Weight of one cubic foot of steam 
at stated pressure. 



100 STEAM ENGINEERING 

PROPERTIES OF SATURATED STEAM 



p 


T 


W 


L 


S 


C 


14.7 


212.0 


180. £ 


965.71 1.146.6 


.03794 


15 


213.0 


181. c 


965.0 


1,146.9 


.03868 


16 


216.3 


185. S 


962.7 


1,147.9 


.04110 


17 


219.4 


188.4 


960.5 


1,148.9 


.04352 


18 


222.4 


191.4 


958.3 


1,149.8 


.04592 


19 


225.2 


194.3 


956.3 


1,150.6 


.04831 


20 


227.9 


197. C 


954.4 


1,151.5 


.0507C 


21 


230.5 


199.7 


952.6 


1,152.2 


.05308 


22 


233.0 


202.2 


950.8 


1,153.0 


.05545 


23 


235.4 


204.7 


949.1 


1,153.7 


.05782 


24 


237.8 


207. C 


947.4 


1,154.5 


.06018 


25 


240.0 


209.3 


945.8 


1,155.1 


.06253 


26 


242.2 


211.5 


944.3 


1,155.8 


.06487 


27 


244.3 


213.7 


942.8 


1,156.4 


.06721 


28 


246.3 


215.7 


941.3 


1,157.1 


.06955 


29 


248.3 


217.8 


939.9 


1,157.7 


.07188 


30 


250.2 


219.7 


938.9 


1.158.3 


.07420 


31 


252.1 


221.6 


937.2 


1,158.8 


.07652 


32 


254.0 


223.5 


935.9 


1,159.4 


.07884 


33 


255.7 


225.3 


934.6 


1,159.9 


.08115 


34 


257.5 


227.1 


933.4 


1,160.5 


.08346 


35 


259.2 


22'8.8 


932.2 


1,161.0 


.08576 


36 


260.8 


230.5 


931.0 


1,161.5 


.08806 


37 


262.5 


232.1 


929.8 


1,162.0 


.09035 


38 


264.0 


233.8 


928.7 


1,162.5 


.09264 


39 


265.6 


235.4 


927.6 


1,162.9 


.09493 


40 


267.1 


236.9 


926.5 


1,163.4 


.09721 


41 


268.6 


238.5 


925.4 


1,163.9 


.09949 


42 


270.1 


240.0 


924.4 


1,164.3 


.1018 


43 


271.5 


241.4 


923.3 


1,164.7 


.1040 


44 


272.9 


242.9 


922.3 


1,165.2 


.1063 


45 


274.3 


244.3 


921.3 


1,165.6 


.1086 


46 


275.7 


245.7 


920.4 


1,166.0 


.1108 


47 


277.0 


247.0 


919.4 


1,166.4 


.1131 


48 


278.3 


248.4 


918.5 


1,166.8 


.1153 


49 


279.6 


249.7 


917.5 


1,167.2 


.1176 


50 


280.9 


251.0 


916.6 


1,167.6 


.1198 


51 


282.1 


252.2 915.71 


1,168.0 


.1221 


52 


283.3 


253.5 


914.9 


1,168.4 


.1243 


53 


284.5 


254.7 


914.0 


1,168.7 


.1266 


54 


285.7 


256.0 


913.1 


1,169.1 


.1288 


55 


286.9 


257.2 


912.3 


1,169.4 


.1311 


56 


288.1 


258.3 


911.5 


1,169.8 


.1333 


57 


289.1 


259.5 


910.6 


1,170.1 


.1355 


58. 


290. a 


260.7 


909.8 


1,170.5 


.1377 


59 


291.4 


261.8 


909.0 


1,170.8 


..1400 


60 


292.5 


262.9 


908.2 


1,171.2 


.1422 



101 



Properties of Saturated Steam — Continued 



! p 


T 


W 


L 


s 


C 


61 


293.6 


264.0 


907.5 


1,171.5 


.1444 


62 


294.7 


265.1 


906.7 


1,171.8 


.1466 


63 


295.7 


266.2 


905.9 


1,172.1 


.1488 


64 


296.8 


267.2 


905.2 


1,172.4 


.1511 


65 


297.8 


268.3 


904.5 


1,172.8 


.1533 


66 


298.8 


269.3 


903.7 


1,173.1 


.1555 


67 


299.8 


270.4 


903.0 


1,173.4 


.1577 


68 


300.8 


271.4 


902.3 


1,173.7 


.1599 


69 


301.8 


272.4 


901.6 


1,174.0 


.162] 


70 


302.7 


273.4 


900.9 


1,174.3 


.1643 


71 


303.7 


274.4 


900.2 


1,174.6 


.1665 


72 


304.6 


275.3 


899.5 


1,174.8 


.1687 


73 


305.6 


276.3 


898.9 


1,175.1 


.1709 


74 


306.5 


277.2 


898.2 


1,175.4 


.1731 


75 


307.4 


278.2 


897.5 


1,175.7 


.1753 


76 


308.3 


279.1 


896.9 


1,176.0 


.1775 


77 


309.2 


280.0 


896.2 


1,176.2 


.1797 


78 


310.1 


280.9 


895.6 


1,176.5 


.1819 


79 


310.9 


281.8 


895.0 


1,176.8 


.1840 


80 


311.8 


282.7 


894.3 


1,177.0 


.1862 


81 


312.7 


283.6 


893.7 


1,177.3 


.1884 


82 


313.5 


284.5 


893.1 


1,177.6 


.1906 


83 


314.4 


285.3 


892.5 


1.177.8 


.1928 


84 


315.2 


286.2 


891.9 


1,178.1 


.1950 


85 


316.0 


287.0 


891.3 


1,178.3 


.1971 


86 


316.8 


287.9 


890.7 


1,178.6 


.1993 


87 


317.7 


288.7 


890.1 


1,178.8 


.2015 


88 


318.5 


289.5 


889.5 


1,179.1 


.2036 


89 


319.3 


290.4 


888.9 


1,179.3 


.2058 


90 


320.0 


291.2 


888.4 


1,179.6 


.2080 


91 


320.8 


292.0 


887.8 


1,179.8 


.2102 


92 


321.6 


292.8 


887.2 


1,180.0 


.2123 


93 


322.4 


293.6 


886.7 


1,180.3 


.2145 


94 


323.1 


294.4 


886.1 


1,180.5 


.2166 


95 


323.9 


295.1 


885.6 


1,180.7 


.2188 


96 


324.6 


295.9 


885.0 


1,181.0 


.2210 


97 


325.4 


296.7 


884.5 


1,181.2 


.2231 


98 


326.1 


297.4 


884.0 


1,181.4 


.2253 


99 


326.8 


298.2 


883.4 


1.181.6 


.2274 


100 


327.6 


298.9 


882.9 


1,181.8 


.2296 


101 


328.3 


299.7 


882.4 


1,182.1 


.2317 


102 


329.0 


300.4 


881.9 


1,182.3 


.2339 


103 


329.7 


301.1 


881.4 


1,182.5 


.2360 


104 


330.4 


301.9 




1,182.7 


.2382 


105 


331.1 


302.6 


88013 


1,182.9 


.2403 


106 


331.8 


303.3 


879.8 


1,183.1 


.2425 



102 STEAM ENGINEERING 

Properties of Saturated Steam — Continued 



p 


T 


W 


L 


S 


C 


107 


332.5 


304.0 


879.3 


1,183.4 


.2446 


108 


333.2 


304.7 


878.8 


1,183.6 


.2467 


109 


333.9 


305.4 


878.3 


1,183.8 


.2489 


110 


334.5 


306.1 


877.9 


1,184.0 


.2510 


111 


335.2 


306.8 


877.4 


1,184.2 


.2531 


112 


335.9 


307.5 


876.9 


1,184.4 


.2553 


113 


336.5 


308.2 


876.4 


1,184.6 


.2574 


114 


337.2 


308.8 


875.9 


1,184.8 


.2596 


115 


337.8 


309.5 


875.5 


1,185.0 
i,185.2 


.2617 


116 


338.5 


310.2 


875.0 


.2638 


117 


339.1 


310.8 


874.5 


1,185.4 


.2660 


118 


33.9.7 


311.5 


874.1 


1,185.6 


.2681 


119 


340.4 


312.1 


873.6 


1,185.8 


.2703 


120 


341.0 


312.8 


873.2 


1,185.9 


.2724- 


121 


341.6 


313.4 


872.7 


1,186.1 


.2745 


122 


342.2 


314.1 


872.3 


1,186.3 


.2766 


123 


342.9 


314.7 


871.8 


1,186.5 


.2788 


,124 


343.5 


315.3 


871.4 


1,186.7 


.2809 


125- 


344.1 


316.0 


870.9 


1.186.9 


.2830 


126 


344.7 


316.6 


870.5 


1,187.1 


.2851 


127 


345.3 


317.2 


870.0 


1,187.3 


.2872 


128 


345.9 


317.8 


869.6 


1,187.4 


.2894 


129 


346.5 


318.4 


869.2 


1,187.6 


.2915 


130 


347.1 


319.1 


868.7 


1,187.8 


.2936 


131 


347.6 


319.7 


868.3 


1,188.0 


.2957 


132 


348.2 


320.3 


867.9 


1,188.2 


.2978 


133 


348.8 


320.8 


867.5 


1,188.3 


.3000 


134 


349.4 


321.5 


867.0 


1,188.5 


.3021 


135 


350.0 


322.1 


868.6 


1,188.7 


.3042 


136 


350.5 


322.6 


866.2 


1,188.9 


.3063 


137 


351.1 


323.2 


865.8 


1,189.0 


.3084 


138 


351.8 


323.8 


865.4 


1,189.2 


.3105 


139 


352.2 


324.4 


865.0 


1,189.4 


.3126 


140 


352.8 


325.0 


864.6 


1,189.5 


.3147 


141 


353.3 


325.5 


864.2 


1,189.7 


.3169 


142 


353.9 


326.1 


863.8 


1,189.9 


.3190 


143 


354.4 


326.7 


863.4 


1,190.0 


.3211 


144 


355.0 


327.2 


863.0 


1,190.2 


.3232 


145 


355.5 


327.8 


862.6 


1,190.4 


.3253 


146 


356.0 


328.4 


862.2 


1,190.5 


.3274 


147 


356.6 


328.9 


861.8 


1,190.7 


.3295 


148 


357.1 


329.5 


861.4 


1,190.9 


.3316 


149 


357.6 


330.0 


861.0 


1,191.0 


.3337 


150 


358.2 


330.6 


860.6 


1,191.2 


.3358 


151 


358.7 


331.1 


860.2 


1,191.3 


.3379 


152 


359.2 


331.6 


859.9 


1,191.5 


.3400 


153 


359.7 


332.2 


859.5 


1,191.7 


.3421 



BOILERS 103 

Properties of Saturated Steam — Continued 



p 


T 


W 


L 


s 


C 


154 


360.2 


332.7 


859.1 


1,191.8 


.3442 


155 


360.7 


333.2 


868.7 


1,192.0 


.3463 


156 


361,3 


333.8 


858.4 


1,192.1 


.3483 


157 


361.8 


334.3 


858.0 


1,192.3 


.3504 


158 


362.3 


334.8 


857.6 


1,192.4 


.3525 


159 


362.8 


335.3 


857.2 


1,192.6 


.3546 


160 


363.3 


335.9 


856.9 


1,192.7 


.3567 


161 


363.8 


336.4 


856.5 


1,192.9 


.3588 


162 


364.3 


336.9 


856.1 


1,193.0 


.3609 


163 


364.8 


337.4 


855.8 


1,193.2 


.3630 


164 


365.3 


337.9 


855.4 


1,193.3 


.3650 


165 


365.7 


338.4 


855.1 


1,193.5 


.3671 


166 


366.2 


338.9 


854.7 


1,193.6 


.3692 


167 


366.7 


339.4 


854.4 


1,193.B 


.3713 


168 


367.2 


339.9 


854.0 


1,193.9 


.3731 


169 


367.7 


340.4 


853.6 


1,194.1 


.3754 


170 


368.2 


340.9 


853.3 


1,194.2 


.3775 


171 


368.6 


341.4 


852.9 


1,194.4 


.3796 


172 


369.1 


341.9 


852.6 


1,194.5 


.3817 


173 


369.6 


342.4 


852.3 


1,194.7 


.3838 


174 


370.0 


342.9 


851.9 


1.194.8 


.3858 


175 


370.5 


343.4 


851.6 


1,194.9 


.3879 


176 


371.0 


343.9 


851.2 


1,195.1 


.3900 


177 


371.4 


344.3 


850.9 


1,195.2 


.3921 


178 


371.9 


344.8 


850.5 


1,195.4 


.3942 


179 


372.4 


345.3 


850.2 


1,195.5 


.3962 


180 


372.8 


345.8 


849.9 


1,195.7 


.3983 


181 


373.3 


346.3 


849.5 


1,195.8 


.4004 


182 


373.7 


346.7 


849.2 


1,195.9 


.4025 


183 


374.2 


347.2 


848.9 


1,196.1 


.4046 


184 


374.6 


347.7 


848.5 


1,196.2 


.4066 


185 


375.1 


348.1 


848.2 


1,196.3 


.4087 


186 


375.5 


348.6 


847.9 


1,196.5 


.4108 


187 


375.9 


349.1 


847.6 


1,196.6 


.4129 


188 


376.4 


349.5 


847.2 


1,196.7 


,4150 


189 


376.9 


350.0 


846.9 


1,196.9 


.4170 


190 


377.3 


350.4 


846.6 


1,197.0 


.4190 


191 


377.7 


350.9 


846.3 


1.197.1 


.4212 


192 


378.2 


351.3 


845.9 


1,197.3 


.4233 


193 


378.6 


351.8 


845.6 


1,197.4 


.4254 


194 


379.0 


352.2 


845.3 


1,197.5 


.4275 


195 


379.5 


352.7 


845.0 


1,197.7 


.4296 


196 


380.0 


353.1 


844.7 


1,197.8 


.4317 


197 


380.3 


353.6 


844.4 


1,197.9 


.4337 


198 


380.7 


354.0 


844.1 


1.198.1 


.4358 


199 


381.2 


354.4 


^ 843.7 


1,198.2 


.4379 


200 


381.6 


354.9 


843.4 


1,198.3 


.4400 















104 STEAM ENGINEERING 

WATER EVAPORATED UNDER WORKING 
CONDITIONS 

When water that is used in conducting 
a boiler test is measured in barrels, it is 
only necessary to multiply the number 
of barrel fulls by the weight of each to 
secure the total weight, but if a meter is 
used, the quantity that has passed 
through it is indicated in cubic feet. 
Care must be taken to note the temper- 
ature as that determines the weight per 
cubic foot. 

Suppose that during a test lasting 10 
hours, the meter indicates that 658 
cubic feet have passed, or 65.8 per hour, 
and calibration of the meter according 
to directions already given shows that 
.98 of its indications are the true quan- 
tity. Then 65.8 X. 98 =64.484 cubic 
feet. Taking the temperature at 65 de- 
grees, the weight per cubic foot is 62.34 
pounds, or 64.484X62.34 = 4,019.93 
pounds per hour. 

For a simple illustration it is assumed 
that after the feed water passes through 
the meter at 65 degrees it goes to a heat- 
er where its temperature is raised to 100 
degrees Fah., and it is then forced into 
the boiler which carries 70 pounds pres- 
sure by the gauge or 85 pounds absolute. 
The atmospheric pressure is 14.7 pounds 
or less, but inasmuch as our ordinary 



BOILERS 105 

steam gauges do not designate fractions 
of a pound it is not necessary to take 
them into account in this calculation. 

Under these conditions every 30 
pounds of water pumped into the boiler 
represents one horse power, therefore 
this boiler developed 4,019.93-^30 = 
134 horse power. 

Particular attention is called to the 
fact that no mention is made of the pur- 
pose for which this steam is used, be- 
cause it makes no difference in the cal- 
culations. Some of it may be used to 
run an engine, another portion to oper- 
ate pumps, and the remainder to heat 
dry kilns, or for any other purpose for 
which steam is required, but this has no 
effect on the power developed by the 
boiler. 

This does not necessarily mean that 
it wiU supply enough steam to run an 
engine and develop 134 horse power, 
because it might require 50 pounds of 
steam per hour for each horse power 
developed. This steam would then fur- 
nish 4,019.93^50 = 80.4 horse power. 

On the other hand, it might be used to 
run a high grade engine requiring only 
15 pounds of steam per hour for each 
horse power developed. Under this 



106 STEAM ENGINEERING 

condition the boiler would supply-:* 
4,019.93 -^15 =268 horse power. 

These two examples clearly illustrate 
the injustice of rating a boiler by the 
power developed in an engine. In the 
former case the engine might be rated 
at 134 horse power, still the boiler could 
not supply the steam required to run it, 
although it might be forced much be- 
yond its rated capacity in an effort to 
keep the engine in operation. On this 
basis the boiler would be condemned, 
although developing more power than it 
was designed for. 

In the latter case the engine might be 
rated at 134 hbrse power and the boiler 
might be highly commended because it 
supplies the required steam with a slow 
fire in the furnace and a small amount 
of coal per hour, but this would not be a 
fair decision, because the boiler would 
be supplying only 134X15-^30 = 67 
horse power which fully explains its easy 
performance. 

ACCOUNTING FOR MOISTURE IN STEAM 

Frequently there is moisture in steam 
supplied by a boiler, therefore, it be- 
comes necessary to take this into ac- 



BOILERS 107 

count in order that the boiler may be 
given credit for its exact performance. 
In this case 4,019.93 pounds of water 
were pumped into the boiler. If the 
calorimeter test shows that 3 per cent, 
of this water is not evaporated, but 
passes out as hot water mixed with the 
steam, it must be subtracted from the 
total weight used. Then 4,019.93- 
(4,019.93 X .03) = 4,019.93 - 120.59 = 
3,899.34 pounds. 

This water went into the boiler at a 
temperature of 100 degrees and passed 
out at 85 pounds absolute pressure, the 
corresponding temperature of which is 
316 degrees F^h. and this heat must be 
accounted for. Water at 100 degrees 
contains 68.08 heat units per pound, 
which is increased to 287 at 85 pounds 
absolute pressure, therefore 287—68.08 
= 218.92 heat units were put into each 
pound of it, or 120.59X218.92=26,- 
399.56 heat units for the whole. 

In order to reduce this to proper terms 
it is necessary to ascertain how many 
pounds of water this amount of heat 
will evaporate under given conditions. 
The total heat of steam at 85 pounds 
absolute pressure is 1,178.3. (See table.) 
It already contains 68.08 heat units per 



108 STEAM ENGINEERING 

pound, therefore, it requires 1,178.3 — : 
68.08 = 1,110.22 heat units. Then 
26,399.56 will evaporate 26,399.56^ 
1,110.22=23.78 pounds of water, if it 
was utilized for this purpose. 

Adding this to the amount actually 
evaporated shows that if all of the heat 
accounted for had been used to convert 
water into steam, the amount would 
have been 3,899.34+23.78 = 3,923.12 
pounds. Dividing this by » 30 shows 
under these conditions the boiler would 
develop 130.77 horse power. 

EQUIVALENT HORSE POWER 

The above title is used because it ex- 
presses concisely the meaning of this 
paragraph when fully explained. Boil- 
ers are seldom operated under the exact 
conditions laid down for standard tests 
when the horse power developed is to be 
determined, but fortunately it is not 
necessary to comply with these condi- 
tions so far as steam pressure carried 
and temperature of feed water are con- 
cerned, as it is not difficult to reduce 
the results secured under any given 
conditions to terms that will admit of 
comparison on a common basis with the 



BOILERS 109 

standard, which is the evaporation of 30 
pounds of water per hour, when carrying 
70 pouaids gauge, or 85 pounds absolute 
pressure, with feed water at 100 degrees 
Fah. ^ 

Every pound (in weight) of steam at 
this pressure contains 1,178.3 heat units, 
and every pound of feed water at 100 
degrees contains 68.08 heat units, there- 
fore heat from the furnace must supply 
1,178.3-68.08 = 1,110.22 heat units. 
This demonstrates that the generation 
and application to water in a boiler of 
enough heat to evaporate 30 pounds in 
one hour where each pound requires 
1,110.22 heat units, constitutes a boiler 
horse power. 

It naturally follows that if steam is 
generated under conditions that require 
less heat per pound, a greater weight of 
water will be evaporated by the same 
amount of heat. Consequently if con- 
ditions are such that more heat is re- 
quired per pound of steam, less water 
will be evaporated by the same quantity 
of heat. 

The quantity of water required per 
hour to constitute one horse power 
under different conditions can be deter- 
mined by the following formula: 






110 STEAM ENGINEERING 

1,110.22 

-^-^X30 = W. 

T = Total heat of steam at given 

pressure. 
F=Heat units in the feed water 

above 32 degrees, at given 

temperature. 
W= Weight of water in pounds. 

For an illustration of the application 
of this formula, suppose that a boiler 
evaporates into dry steam 4,019.93 
pounds of water per hour, under 165 
pounds absolute pressure, with feed 
water at 210 degrees Fah. How much 
water constitutes one horse power under 
these conditions and how much power 
is developed? 

The total heat of steam at 165 abso- 
lute, or 150 pounds gauge pressure is 
1,193.5 and water at 210 degrees con- 
tains 178.87 heat units 

1,110.22 

X30 = 32.826 pounds of 

1,193.5-178.87 ^ 

water required to constitute one horse 

power. 4,019.92^32,826 = 122.46 horse 

power developed. 

As the evaporation of the same weight 

of water under less favorable conditions 



BOILERS 111 

developed 134 horse power, the im- 
proved conditions, consisting of heating 
the feed water to a higher temperature, 
reduced the load by 11.54 horse power. 
This reduction of load results in a 
corresponding saving in fuel, calling at- 
tention to great benefits derived from 
heating the feed water by exhaust steam 
which is frequently a waste product. 

The next table gives the weight, in 
pounds, of water that must be evapor- 
ated per hour to constitute one horse 
power under different conditions. It 
is based on the foregoing formula, and 
will be very useful to engineers and 
others who wish to know at a glance 
how much water is required for one 
horse power under conditions found in 
their respective plants. 

For illustration, suppose that a certain 
boiler is operated under 125 pounds gauge 
pressure and the feed water enters at 
210 degrees after passing through any 
kind of a heater that utilizes exhaust 
steam. Following the column under 
140 pounds absolute pressure (which is 
equal to 125 by the gauge) until it in- 
tersects with the line beginning with 
210 it shows that 32.95 pounds are re- 
quired under these conditions. 



112 



STEAM ENGINEERING 



POUNDS OP WATER PER HORSE POWER 

Absolute Boiler Pressure 



40 
50 
60 
70 
80 
90 
100 
110 
120 
130 
140 
150 
160 
170 
180 
190 
200 
210 
212 



28.74 
28.97 
29.23 
29.48 
29.75 
30.02 
30.29 
30.57 
30.85 
31.14 
31.44 
31.74 
32.05 
32.36 
32.68 
33.01 
33.34 
33.69 
33.75 



28.56 

28.80 

29.05 

29.31 

29.57 

29.84 

30.11 

30.39 

30 

30.95 

31.24 

31.54 

31.84 

32.15 

32.46 

32.79 

33.12 

33.46 33 

33.53133 



28.37 
28.61 
28.86 
29.12 
29.37 
29.63 
29.90 
30.17 
30.45 
30.73 
31.02 
31.32 
31.61 
31.92 
32.23 
32.55 
32.87 
33.21 
33.27 



POUNDS OF WATER PER HORSE POWER 
Absolute Boiler Pressure 



40 

50 

60 

70 

80 

90 

100 

110 

"120 

130 

140 

150 

160 

170 

180 

190 

200 

210 

212 



28.08 
28.32 
28.56 
28.80 
29.08 
29.31 
29.57 
29.84 
30.11 
30.39 
30.67 
30.96 
31.25 
31.55 
31.85 
32.16 
32.48 
32.80 
32.87 



BOILERS 113 

BENEFITS OF FEED WATER HEATERS 

Although the saving in fuel is more 
than enough benefit to pay for the in- 
stallation of a good feed water heater, 
it is not the only advantage gained by 
its use, and under some conditions the 
saving of unnecessary strains on the 
boiler is a greater compensation than 
the saving in fuel. 

This is especially true of locomotive 
boilers or any other type that is fitted 
with a water leg into which the cool feed 
water is discharged. The Hartford 
Steam Boiler Inspection & Insurance Co. 
made experiments some time ago which 
demonstrated that cool water coming in 
contact with a heated boiler plate causes 
it to contract, and as such a plate is 
rigidly connected to others, a very great 
strain on these parts is the sure result. 

While it is difficult or impossible to 
calculate with even a reasonable degree 
of accuracy what these strains amount 
to, it is a well known fact that they cause 
cracks in the plates which greatly weaken 
them, and if such cracks are not given 
intelligent attention, the result may be 
disastrous explosions. 

As plates contract directly in propor- 
tion to the difference between the tem- 
perature of the feed water striking 
directly against them, and that of the 
steam or hot water in contact with ad- 



114 STEAM ENGINEERING 

jacent parts, anything that reduces this 
difference cannot fail to be beneficial. 
The girth seams of horizontal boilers 
frequently leak from the same cause. 

The percentage of fuel saved by the 
installation of a good feed water heater 
that utilizes exhaust steam, or by an 
economizer that takes heat from waste 
gases on their way to the chimney, can 
be determined by the following rule. 

From the total heat, or heat units in 
the heated water, subtract the heat 
units in the cold water. Divide the re- 
mainder by the total heat of steam at 
the pressure carried, minus the heat 
units in the cold water. Multiply the 
quotient by 100. 

When written as a formula it appears 
as follows: 

|^^X100 = P, 

H = Total heat in the heated water. 

C = Total heat in the cold water. 

S = Total heat in the steam at the 

pressure carried, 
P = Percentage of gain, or the por- 
tion of fuel saved by heating 
the feed water. 
For illustration, suppose that a boiler 
carries 140 pounds absolute pressure and 
uses water from the street main at a 
temperature of 50 degrees Fah. What 
percentage of the fuel now burned will be 



BOILERS 115 

saved by the installation of an exhaust 

steam feed water heater that will raise 

the temperature of it to 211 degrees 

Fah., assuming that the exhaust steam is 

not utilized for any other purpose? 

Applying the formula to this case 

results as follows: 

179.89-18 

X 100 = 13.8 per cent. 

1,189.5-18 

It is not a difficult matter to determine 
the saving that will result under any 
conditions that can be found in practice 
by the foregoing rule and formula, using 
tables of properties of water and steam 
that are found on preceding pages, but 
for the convenience of readers the next 
table contains the results of calculation 
to determine these values. The initial 
temperatures refer to the temperature 
of water as it enters the heater. This 
is given from 40 to 100 degrees Fah. as 
this range will be sufficient to cover alT 
ordinary cases. For lower or higher 
temperatures use the formula. 

The table includes cases where, the 
water is heated to 180, 190, 200, and 210 
degrees Fah., as a heater that will not 
deliver water at the former temperature 
is of little value, and the latter is seldom 
exceeded without back pressure on the 
engine. The pressures are absolute, 
therefore 15 must be subtracted from 



116 



STEAM ENGINEERING 



them to secure gauge pressures. The re- 
sults given are the percentage of saving 
on all fuel burned where water is forced 
into boilers at the initial temperatures. 





o 


11.85 
11.10 
10.33 
9.56 
8.77 
7.96 
7.14 




§ 


11.86 
11.11 
10.34 
9.57 
8.78 
7.97 
7.15 




g 


11.87 
11.12 
10.35 
9.58 
8.79 
7.98 
7.16 




o 


11.89 
11.14 
10.38 
9.59 
8.80 
7.99 
7.17 


Pi 

t3 


"-^ 


11.91 
11.16 
10.39 
9.61 
8.82 
8.00 
7.18 




11.93 
11.18 
10.41 
9.63 
8.83 
8.02 
7.19 


Pi 


o 


11.95 
11.19 
10.42 
9.64 
8.84 
8.03 
7.20 


oi 
w 
h4 


s 


11.97 
11.21 
10.44 
9.66 
8.86 
8.05 
7.22 


g 

i 


§ 


11.99 
11.24 
10.46 
9.67 
8.88 
8.06 
7.24 


§ 


•-H CO 03 O O 00 o 
O IM Tt<vl> 03 O (N 

drHddododt>l 


o 


12.05 
11.29 
10.51 
9.73 
8.92 
8.10 
7.27 




s 


12 08 
11.32 
10.54 
9.75 
8.95 
8.13 
7.29 




g 


12.12 
11.35 
10.58 
9.78 
8.97 
8.15 
7.31 




^ 


12.16 
11.39 
10.61 
9.82 
9.01 
8.18 
7.35 






5g8g§§S 



(Nr-lr-lO03 00 00 



rH t>. r-( rt< CD CD iO 

r^qc^Tiiqoqo 
im" r-5 .-5 d d 00 00 



(N --I rH O 05 00 00 



<NC^»-(0 03 00 00 



. t^lNCOOOOOO 

I i>;q<N^i>qq 

I y c4(NrHddo6o6 

I Ph 00 tJ< 00 .-I rH C . _. 

, p t^OC<HOt>030 

; w c<i(N.-Hddo6o6 

; W O CO O iM CO CO i-H 

^ p^ 00 q CO >o r- Oi »H 

I pu, (NNi-nddodoo 



I J C<1(M r-( O 03 00 00 



(N<Nr-tOO>00 00 






KNr-iOOlOOOO 



N(N I-H 00 03 00 



<N (M 1-1 03 03 00 

03cocD'r^oococo 
q<N ■* q 00 o (N 
cq'ci T-5dddo6 



■*00 OC<lT-iOit» 

qcN »o t> C3 q oj 
eoc<Jr-idddo6 



o .5 ooooooo 

O > rt>Ui<Ot»OOaiO 



117 





o 


13.55 
12.81 
12.06 
11.31 
10.53 
9.74 
8.94 




s 
I 

o 


13.56 
12.83 
12.08 
11.32 
10.5-4 
9.76 
8.95 




13.58 
12 85 
12.09 
11.33 
10.56 
9.77 
8.96 




13.60 
12.86 
12.11 
11.35 
10.57 
9.78 
8.97 


Pi 
a* 

oi 


§ 


13.62 
12.88 
12.13 
11.37 
10.59 
9.79 
8.99 


§ 


13.65 
12.90 
12.15 
11.39 
10.61 
9.81 
9.00 


o 


13.67 
12.93 
12.17 
11.40 
10.62 
9.83 
9.02 


o 


1 


13.69 
12.95 
12.19 
11.43 
10.64 
9.85 
9.03 


p-i 

H 


§ 


13.72 
12.98 
12.22 
11.45 
10.66 

9!05 


O 


§ 


cocoN'-'doJd 


< 


s 


13.78 
13.04 
12.28 
11.50 
10.72 
9.91 
9.10 




8 


13.82 
13.07 
12.31 
11.53 
10.75 
9.94 
9.12 




o 


13.86 
13.11 
12.35 
11.57 
10.78 
9.98 
9.14 




s 


13.91 
13.16 
12.39 
11.61 
10.82 
10.01 
9.19 




IP 


5SSg§§8 



Tl<CO(N(N'-i0 05 



CO(N(Nr-tOOi 



4> ;::<;:; ;;HrHi.H 



■* CO <M <N r-l O 0> 

COCCOOCOOOOO 
■* >; 05(N -"l^ CO 

-^-eoNdt-Hdc 



= JO Oi- 



j ^COCONt-iOOS 



OCONrHOOS 



a «3 <NO5iOO5(MC0C0 

fe K iqt>;qN"5t^q 

Z. ^ TticOCONrHdcJ 



S P$ COtNOO '-H'^'OCO 

•g fq ioooqcoiqt-;q 
- d ■^coco<N'-Hdd 

r O 00 lo o •* CO t> i>. 
^ P5 40oqrHco"5i>q 
o la '^Mco(N-'-5dd 

^ f:' (M 00 CO CO 00 05 OS 
^^ tD ;O00rHC0»Ot>;O5 

^ h4 -^'cococQj^dd 

> W t> rH CO Oi -"-I --H 

P* pq qqi-HcoqooQ 
•a <J -"i^ CO CO (m';^ do 



■^ t>-q(Nn<qoqq 
'o -^ CO CO 'N 'H d d 



CO Oi ■* t> 00 00 00 
t- q (N '^ CO 00 q 

■* CO CO <N »-H d d 



*< ■* 00 (N >-i O O 



EQUIVALENT EVAPORATION 

When a certain weight of water is 
evaporated in a steam boiler at a known 



118 STEAM ENGINEERING 

pressure, it forms a basis for computing 
the power developed, as before explained, 
also for comparing results secured with 
other boilers, but for convenience and 
accuracy it is customary to reduce all of 
them to a common level. 

Boilers are fed with water whose tem- 
perature varies through a wide range, 
and steam is generated under many 
different pressures, but when it is as- 
sumed that the feed water is heated to 
212 degrees Fah. and that steam is gen- 
erated under atmospheric pressure at a 
temperature of 212 degrees Fah. and all 
results are reduced to these standards, 
comparisons can be intelligently made 
at short notice. 

As the feed water and the steam gen- 
erated have the same temperature, it 
demonstrates that only the latent heat 
of steam at atmospheric pressure is put 
into the water after it enters the boiler. 
This amounts to 965.7 heat units per 
pound above 32 degrees Fah., and this 
result is obtained by subtracting the 
heat units in a pound of the feed water 
at this temperature, from the total heat 
of steam at this pressure, and 1,146.6 — 
180.9 = 965.7. If water at 100 degrees 
is pumped into a boiler carrying 85 
pounds absolute pressure, each pound of it 
requires 1,178.3-68.08 = 1,110.22 heat 
units to evaporate it into dry steam. 



BOILERS 119 

Thus it becomes plain that the numbers 
965.7 and 1,110.22 are used for compari- 
sons along this line. 

The evaporation of 30 pounds of water 
in one hour, where each pound of it re- 
quires 1,110.22 heat imits, takes 33,306.6 
heat units. Under other conditions 
the evaporation of one pound of water 
requires 965.7 heat units, therefore the 
total just mentioned will evaporate 
33,306.6-^965.7 = 34.489 pounds of 
water, which is called 34.5 for conve- 
nience in calculating results. Thus the 
evaporation of 34.5 pounds of water in 
one hour from and at 212 degrees con- 
stitutes one boiler horse power, as it is 
equivalent to the standard explained on 
foregoing pages. 

Another way in which this can be used 
is to show how much water would have 
been evaporated in any given case found 
in practice, provided the feed water was 
at a temperature of 212 degrees and 
there was no pressure by the gauge. 
Such a problem is solved by the fol- 
lowing rule: 

Multiply the weight of water actually 
evaporated into dry steam, by the num- 
ber of heat units required to evaporate 
one pound and divide the product by 
965.7. The quotient is the weight that 
would be evaporated under assumed 
conditions. 



120 STEAM ENGINEERING 

For example, suppose that a boiler 
evaporates into dry steam 4,019.93 
pounds of water in one hour, from feed 
water at 100 degrees, into steam at 70 
pounds gauge pressure. How much 
would have been evaporated from and 
at 212 degrees? How much power 
would have been developed? 

4,019.93 X (1,178.3 - 68.08) -^ 965.7 = 
4,621.5 pounds. 

When this is divided by 34.5 it shows 
that 134 horse power was developed. 

COAL REQUIRED TO EVAPORATE WATER 

Owners of steam plants and superin- 
tendents of mills, shops and factories fre- 
quently say that if two kinds of coal are 
to be tested it is only necessary to com- 
pare the weight used to run the machin- 
ery one week, with what is required to 
operate it another week, and that 
comparison decides the comparative 
value of two kinds of coal. 

To a certain limited extent this con- 
clusion is correct, because the weight 
of coal burned, in connection with the 
price per ton determines the cost of fuel 
for a given time, but it is unjust and 
unfair to base an opinion concerning 
the real merit of any kind of fuel on the 
result of such a so-called test, which is 



BOILERS 121 

SO crude and unsatisfactory that it is 
not worthy to be called a test at all. 

When discussing the merits of such a 
transaction the steam user always claims 
that the same amount of machinery was 
operated during the time mentioned, 
but while such a claim may seem rea- 
sonable from his point of view, it really 
is not, because an engine will seldom or 
never develop just the same power for 
many days in succession, even if the 
machines operated are engaged on a 
class of work that apparently does not 
vary, and for some other kinds the 
variation is great from day to day. 
Where live steam is used in varying 
quantities there is much more chance 
for a difference in actual results. 

The quality of the steam produced may 
not change, but this is not known defin- 
itely unless careful tests are made of it 
in accordance with instructions fo\ind 
on foregoing pages of this work. 

The actual weight of coal consumed 
may not have been correctly reported, 
because some kinds contain much more 
moisture than others, and water in the 
furnace should not be counted as coal. 
When a boiler test is to be made, a 
fair sample of the coal should be selected, 
carefully dried and the percentage of 
moisture it contains accurately deter- 
mined. It is usually convenient to take 



122 STEAM ENGINEERING 

a small wooden box, fill it with a sample 
of the coal, then place it on the boiler 
setting to dry. Unless the box has just 
been kiln dried it is sure to contain 
moisture, and as this evaporates when 
the coal is dried, it makes enough differ- 
ence to spoil the effort to fairly rate the 
coal used in making the boiler test. 

In a certain case the small box used 
for this purpose weighed one-half pound 
less after it had been thoroughly dried, 
and the moisture actually evaporated 
in drying the sample of coal was .7 
pound. If the moisture coming out of 
the box had been credited to the coal, 
and the whole pile had been judged by 
the sample, the entire test would have 
been worthless. 

If the sample, as taken from the pile, 
weighed 18.25 pounds and after being 
thoroughly dried it weighed 17.5 pounds 
it shows that 18.25- 17.5 -^ 18.25 X 
100=4.1 per cent, of the weight of ma- 
terial brought from the coal yard was 
water and 95.9 per cent, was coal. By 
courtesy the whole weight is called coal, 
and what is left after the weight of 
moisture is subtracted is called dry 
coal. 

Assuming that 4,019.93 pounds of 
water were pumped into the boiler in 
one hour, and that there was 3 per cent, 
of moisture in the steam produced, the 



BOILERS 123 

total weight evaporated is 3,899.34 
pounds, and the heat used in raising the 
temperature of this moisture is equiva- 
lent to the evaporation of 23.78 pounds, 
then the final weight accounted for is 
3,923.12 pounds. 

If 448 pounds of coal were required 
and it contained 4.1 per cent, of moisture 
which is 18.36 pounds, the actual weight 
of dry coal is 448 - 18.36 = 429.64 
pounds. The actual evaporation is 
therefore 3,923.12-^429.64=9.13 pounds 
of water per pound of dry coal. 

A rule has already been given and 
explained for determining the weight 
of water that would have been evapor- 
ated with feed water at 212 degrees and 
steam at zero by the gauge. In that case 
the total weight of water evaporated per 
hour was taken, but in this case it is the 
weight evaporated per pound of coal 
that is to be compared with the given 
standard. 

Then 9.13 X (1,178.3 - 68.08) ^ 965.7 
= 10.5 pounds from and at 212 degrees. 

WATER EVAPORATED PER POUND OF 

COMBUSTIBLE UNDER GIVEN 

CONDITIONS 

Another point to be taken into con- 
sideration in this connection is that after 
a boiler test is finished, so far as the 



124 STEAM ENGINEERING 

burning of coal is concerned, there is 
more or less ash left which cannot be 
burned. When we consider that this 
varies greatly with different kinds of 
coal, the injustice of charging it to the 
boiler as actually burned, is at once 
plain, therefore it must be taken from 
the weight of dry coal burned. 

Examination of several reports of 
boiler tests shows that the ashes remain- 
ing ranged from 6.4 to 26.2 per cent, of 
the entire weight of dry coal fed into the 
furnace. This shows that it is neces- 
sary in every case to determine the 
weight of dry ashes remaining at the 
conclusion of a test. 

Suppose that in this case there were 
62.75 pounds of ashes left, for each hour 
covered by the test. This would be 
429.64-62.75-^429.64 X 100 = 85.3 per 
cent, of combustible, and as the entire 
weight is represented by 100 the ash 
amounts to 100 — 85.3 = 14.7 per cent. 

Subtracting 62.75 pounds from 429.64 
shows that the weight of combustible is 
366.89 pounds and as this evaporated 
3,923.12 pounds of water, it demon- 
strates that 3,923.12 -=-366.89 = 10.7 
pounds were evaporated for each pound 
of combustible burned under given con- 
ditions. 



BOILERS 125 

WATER EVAPORATED PER POUND OF 

COMBUSTIBLE FROM AND AT 

212 DEGREES FAH. 

While the foregoing example seems 
to be complete, there is another point 
to be disposed of in order to cover the 
whole process. This consists of demon- 
strating the weight of water that would- 
be evaporated per pound of combust- 
ible from and at 212 degrees Fah. 

As 10.7 pounds were evaporated under 
given conditions which required, 1,110.22 
heat units for each pound of water, then 
each pound of coal yielded 1,110.22 X 
10.7 = 11,879.35 heat units. As it 
requires 965.7 to evaporate one pound 
"from and at 212 degrees," this is equiv- 
alent to the evaporation of 12.3 pounds 
under the assumed conditions, accord- 
ing to a rule given and explained on 
previous pages. 

This forms a correct and equitable 
standard for comparison, and it is the 
only one worthy of serious consideration 
where accuracy is desired, as anything 
else that is worked out with less care is 
of less value, and in many cases the 
results secured by incorrect methods 
are worse than useless. 

Where the actual results secured 
from several tests made on different 
boilers, to determine their efficiency in 



126 STEAM ENGINEERING 

service, and these results have all been 
reduced to this common standard, it is 
not necessary to study all of the reports 
as the weight of water evaporated into 
dry steam per pound of combustible, 
"from and at 212 degrees," enables the 
engineer to compare them intelligently 
at short notice. 

THE EFFICIENCY OF STEAM BOILERS 

Steam users are frequently asked to 
buy appliances that will save coal, and 
in hundreds of plants there are oppor- 
tunities to do all that salesmen guaran- 
tee for their goods, but in some cases 
the conditions have not been thoroughly 
investigated, causing enthusiastic ad- 
vocates of certain inventions to make 
extravagant claims for them which can- 
not be realized in practice. 

These remarks do not apply to cases 
where it is proposed to save steam and 
thus save coal indirectly, as this is not 
the proper place to treat them, but only 
to plants where it is expected to save 
coal directly by some appliance for the 
furnace. 

Efforts are made in some of these 
cases to convince steam users that only 
10 or 15 per cent, of the coal burned in a 
boiler furnace is taken up by the water 
when it is converted into steam. If 



BOILERS 127 

this was true there would be a good 
opportunity to save a portion of such a 
great loss, but unfortunately for these 
enthusiasts, and fortunately for the 
steam user, these claims are not true, as 
the following example demonstrates. 

The report of a test made on a verti- 
cal fire tube boiler of the most simple 
kind, shows that for each pound of coal 
burned in the furnace, 11.34 pounds of 
water were evaporated into dry steam 
"from and at 212 degrees." As it re- 
quires 965.7 heat unit's to evaporate one 
pound under these conditions, to evap- 
orate 11.34 pounds it is necessary to use 
10,951 heat units. 

The number of heat units in the brand 
of coal used is not stated, but much of 
our ordinary steam coal does not con- 
tain more than 13,000 heat units per 
pound, in which case the efficiency of 
this boiler under ordinary working con- 
ditions is 10,951 -r- 13,000X100 = 84 per 
cent. Assuming that the coal con- 
tains 13,500 heat units, the efficiency is 
81 per cent, and for 14,000 heat unfts it 
is 78 per cent. This should show every 
reader that it is impossible to save from 
25 to 50 per cent, which has been claimed 
in some cases. 

The vertical fire tube boiler apparent- 
ly affords an excellent chance for heat 
to ascend in the tubes and escape to the 



128 STEAM ENGINEERING 

stack, but this is more apparent than real j 
provided the boiler is well proportioned. , 
During the above mentioned test, the'' 
temperature of the escaping gases was 
427 degrees. The steam pressure was 
60 pounds absolute, the temperature of 
which' is 292.5 degrees, and it was super- 
heated 18 degrees by the tubes passing 
through the steam space, therefore the 
final temperature was 310.5 degrees. 

As the escaping gases were only 116.5 
degrees above this, there cannot be 
much improvement made at this point, 
because there must be more heat in the 
gases than there is in the steam, or else 
heat will escape from the latter to the 
former, thus lowering the efficiency of 
the boiler. 

This is the principal source of loss 
from a well set boiler and an ordinary 
furnace, but measures should be taken 
to prevent radiation of heat from all 
other parts, and all holes in brick setting 
around cast iron fronts, in connection 
with stacks, and in all other places where 
cold air can be drawn in to cool the 
heated surfaces, should be carefully 
closed, as each one represents loss of 
drawing power on the fire, or of heat 
from the coal consumed. This is an im- 
portant point, for when cold air is drawn 
inward the defect is not noticed as 
quickly as when heat comes outward , 



BOILERS 129 

yet all such imperfections are a direct 
source of loss. 

THE LOAD AND THE STEAM PRESSURE 

The load on a steam boiler is frequently 
estimated by the pressure carried on 
it in everyday service, as men who are not 
well informed on the subject of steam 
engineering conclude that when a boiler 
carries a high pressure there must neces- 
sarily be a heavy load on it, but while 
this may be true, and it frequently is, 
still it cannot be laid down as a principle 
because there are many boilers in use 
carrying a low pressure for heating pur- 
poses, yet the load on them is heavy. 

The load on a boiler is determined by 
the weight of water evaporated into dry 
steam in a given time, and if 20 pounds 
gauge pressure is enough for a certain 
place, that does not prevent conditions 
from demanding a rapid rate of evap- 
oration. 

On the other hand the nature of the 
business for which another boiler supplies 
steam may be such that while a gauge 
pressure of 200 pounds or more may be re- 
quired, only a comparatively small quan- 
tity of steam is used, hence the load is light. 

Nobody would allow a boiler used 
under such conditions to be run without 
a fireman or an engineer in constant 



130 STEAM ENGINEERING 

attendance, yet we are sometimes asked 
if steam cannot be left on a building all 
night, without anybody in attendance 
on the boiler, because there is not a very 
high pressure on, consequently there 
can be no danger. 

Suppose a boiler should be left alone 
all night, and soon after the fireman has 
departed the gauge glass should break. 
What would be the consequence by the 
time that he returned in the morning? 
What would become of a fire if left 
without attention for 10 or 12 hours? 

Boilers are frequently run all day, the 
fires banked at night, while the pressure 
is but little lower than is required during 
the day, and when the fireman returns 
the next morning he finds a working 
pressure of steam on. 

This does not prove that steam could 
be used during the night without loss, 
as some managers seem to think, for if 
heat is taken from the boilers after the 
fires are banked there is just so much 
less left for use in the morning. What- 
ever remains at night after a day's run, 
is not lost as it is only stored for future 
use. Coal that is used for banking a fire 
is either there for use when a bright fire 
is wanted or else a portion of it is burned 
and the resulting heat has kept the boil- 
er just so much nearer ready for use 
when wanted at short notice. 



SECTION 2 
BOILER FEEDERS 

REMARKS ON BOILER FEEDERS 

When a boiler feeder is to be selected 
for a steam plant, there are several 
points to be considered, which are ap- 
parently not always remembered, judg- 
ing by the illogical selections made in 
some cases. This does not include the 
efficiency of the boiler feeder as a ma- 
chine, for that will be considered sepa- 
rately, because its importance warrants 
it, but to other conditions which make 
a certain kind suitable for some places 
and unsuitable for others. 

For illustration, take a plant where the 
nearest revolving shaft from which 
power can be obtained is perhaps 100 
feet distant from the boiler room. A 
power pump is not suitable for such a 
plant because the fireman will have to 
leave his boilers too many times in order 
to attend to the pump and regulate the 
supply of feed water delivered, thus not 
only diverting his attention from duties 
in front of his boilers which require con- 
stant care and labor, but also causing 
much useless walking, and furthermore 
such practice results in his being absent 
from the boiler room without good rea- 



132 STEAM ENGINEERING 

sons whenever he feels so inclined, to the 
detriment of good service. 

If more or less of the feed water is 
warm before it passes through a heater 
it is not a good idea to use an injector, as 
it is not reliable under such conditions. 
Many of them will take warm or even 
hot water when new and in perfect order, 
but after they have been used long 
enough to cause slight wear in the tubes 
they "kick" especially if the water is 
warmer than usual, for in all cases the 
incoming water must be cool enough to 
condense the steam used to operate the 
injector. 

It will pay to install a good feed water 
heater in practically every case, but 
where this valuable feature must be 
omitted for any cause, an injector should 
always be used as it delivers hot water 
to the boilers, thus preventing the great 
strains on plates, due to feeding cold 
water. 

In electric power and lighting plants 
where current is always available a 
power pump driven by a motor is con- 
sidered an up-to-date machine. As it is 
practicable to vary the speed of such a 
motor to meet the requirements of one 
or more boilers, and as it may be located 
in the most convenient place without 
regard to the nearest shaft or the most 



BOILER FEEDERS 133 

available steam pipe, it forms a very 
desirable combination for feeding boilers. 

THE EFFICIENCY OF BOILER FEEDERS 

The mechanical efficiency of a boiler 
feeder is found by dividing the power 
actually used in forcing water through' 
pipes and valves into boilers against the 
pressure carried, by the power required 
to operate the machine. In the form of 
a formula it appears as follows : 

A 

' — =M.E. 
P 

A = Actual power required to force 

water into boilers. 

P = Power required to operate the 

machine. 

M E = Mechanical efficiency. 

For example, suppose that 4.8 horse 

power are actually used in forcing water 

into a battery of boilers, and it requires 

6 horse power to operate the machine. 

The mechanical efficiency is therefore 

4.8 

— =.80 
6 

From this example it will be plain that 
the mechanical efficiency of a boiler 
feeder (or any other machine), is always 
a fraction, because it requires a perfect 
machine in which there is absolutely no 



134 STEAM ENGINEERING 

power lost, to equal unity, or 1, and it is 
impossible to devise and construct such 
a machine. 

If a pump is used for this purpose and 
the pistons are packed very tightly, the 
value of P will be large, consequently 
the fraction resulting from an applica- 
tion of the formula is small, or in other 
words the mechanical efficiency is low. 

Where it is desired to express the 
efficiency in the form of a percentage of 
the whole power used the formula be- 
comes 

— X100 = M. E. 

The thermal efficiency of a boiler feed- 
er is the amount of heat used in forcing 
water into the boilers, divided by the 
heat put into the machine to operate it. 
It is found in accordance with the above 
explanation relating to mechanical 
efficiency. 

POWER PUMPS 

When selecting a boiler feeder the 
mechanical efficiency of all kinds taken 
into consideration should be noted, but 
a certain kind may show high mechan- 
ical efficiency when taken as a machine 
designed for this purpose, and when used 
in connection with some other appliance 
the combination may give excellent 



BOILER FEEDERS 135 

results, but when the boiler feeder is 
considered alone it may be wasteful and 
unsatisfactory. 

The improved up-to-date power pump 
is a good illustration of this principle 
because when in good repair its mechan- 
ical" efficiency is high. Furthermore, 
the quantity of steam required to oper- 
ate it is small, as the power needed is 
developed in the main engine which is 
generally economical in the use of steam. 
Still it would be very poor practice to 
use a power pump to force water into, 
one or more boilers without using a heat- 
er to raise the temperature as high as 
possible. This combination gives ex- 
cellent satisfaction in practice, provided 
the pump can be conveniently located 
near a shaft used for other purposes, so 
that but little power will be used in oper- 
ating the transmission devices, what- 
ever they may be, between the main 
engine and the pump. 
Fig. 17 illustrates a single acting 

! power pump fitted with o^e cylinder. 
Water is drawn in on the upward stroke 

; an^ forced out on the downward stroke 
of the plunger. It is a very simple 

! pump, but its capacity is small and its 
mechanical efficiency is comparatively 
low, as there is but one effective stroke 

1 for each revolution of the shaft. 



136 



STEAM ENGINEERING 




FI&./.7 



When a pump of this kind is fitted 
with two cylinders, its capacity is 
doubled and its mechanical efficiency is 



BOILER FEEDERS 



137 



higher, because the friction is not in- 
creased in proportion to the increase in 
capacity. As there are two effective 
strokes for each revolution, the flow of 
water is nearly continuous, although its 
speed is not constant. When fitted 
with three cylinders this is known as the 




FIG. /a 



A 



138 



STEAM ENGINEERING 



single acting triplex pump. It delivers ; 
a nearly constant stream of water. 

Fig. 18 illustrates a double acting ' 
power pump fitted with one cylinder. 
As it draws in water and expels it at 
each stroke, it gives fair results in prac- 
tice. 




FIG, /? 



Ita 



BOILER FEEDERS 139 

When this kind of a pump is fitted 
with two cylinders, and the driving 
cranks are set at an angle of 90 degrees, 
the flow of water is practically continu- 
ous, and the mechanical efficiency is 
satisfactory. 

These pumps are sometimes fitted 
with three cylinders, and as the pistons 
are double acting, an even and contin- 
uous flow of water is secured, when the 
driving cranks are set at an angle of 120 
degrees. 

Fig. 19 illustrates the arrangement of 
valves in this machine, which is known 
as the double acting triplex pump. 

HORSE POWER REQUIRED BY PUMPS 

The first part of the next table gives 
the power required to operate Gourd's 
single acting triplex pumps, when forc- 
ing water against columns of water of 
stated height, or in other words the 
given head in feet, which gives corre- 
sponding pressure. The head includes 
the height from the point where water is 
lifted to the highest part of the delivery 
pipe. 

The second part applies to double 
ing triplex pumps at given capacities. 



140 STEAM ENGINEERING 

When comparing the power required by 
these two kinds of pumps, the capacity 
must be considered in connection with 
the size of pump. 

For illustration, a 7X8 inch single act- 
ing pump when working against 108 
pounds pressure requires 15 horse power 
for the given capacity. A 7X8 inch 
double acting pump would require 30 
horse power at the same rate, but the 
capacity of the * latter is not twice as 
much as the former, therefore when 
the table states that 28 horse power 
will be required, it is a logical con- 
clusion. 

The first column contains the diam- 
eter and stroke of the pump under the 
heading "D, & S." The second gives 
the usual capacity, or the number of 
gallons that each will deliver at normal 
speed under the heading of "Capacity." 
Succeeding columns give first the head 
or height of the column of water, and 
under it the pressure due to this height 
of water. This represents boiler pres- 
sure where the pump is used to feed 
boilers instead of raising the water to 
the given height. Figures under these 
headings are the horse power required 
for stated conditions. 



BOILER FEEDERS 



141 



CO 

8 

CO 

^^ 

G 

a- 

"o 
"o 

h 

8 


1 

d — 

WOO 

(X 

h 

§ 

a 
On 


CO CO -1 (M. O (N. t^ i-O lOO 00 00 O CO lO 


r-i(MC0C0^«300O(NiOt^r>.-<C0i0 
-1 ^ --< ^ --< CO IM (N 


ooooooooooooooo 
(Mocn'-it>50iooiooooooio 


rHl^^(^>cocoo^.C5|^co'Oloa)-^co 


O 00 lO t^ O'OO (N lO t^ O t^ I> O CSI lO 


'-i'riiMC^JCOTt<COt»a)— <(M(NiOt^Cr3 


8§8^^t!^88§^88888 


--•MiMiMCOiOCOt^OOOOiMO-JiO 


g^§§82888888§t28 


T-ir-ir^cqcO'^ocot^ocooasoiM 


g82^gS§8SS888§8 


OQ'OCOOOOOOOLO>OOiOiO 
lOtOt^OOC^JOOiOiOCOOlMOCOOO 


rH,-(lM(NC0CO-*TtiiOiO<O 


a 

c3 


--H -< ^ -< rH (M (N 






CO-*Tt<Tti-:fcOC!COOCOo2o00o2oO 
X X X X X X X X X X X X X X X 

<N(MC0C0-<*-'*'0>0'0OOOt-,I>00 



142 



STEAM ENGINEERING 



o 


o 


o 




CO 






o 




CO 








c 


u 




,■">- 


cr— 


a 




^o 




<uco 


^^ 


ao 




c 

a 


£ 


3 — 


c — 






O 




o 


(^ 




:3r-. 






vi 




o. 


v£i — 


M— 


c; 


C 


r^ 




.Sf^ 


OcD 


ffi^ 


?1- 








o 


o 




o 


'^ 


- 


- 


s 


(N 


>. 








cl 


a 




o 


w 


oa 




Q 



l>0000O(N>0000>i-lC0'*C0t»00'-lTl 



0(N0000e3(N(N(N<©<N-««<e«CSIN 

ooooaiO«-tNooo>o>-ioNwco-<*"> 



BOILER FEEDERS 



143 



i^ 4) 



CD O <M ^ O O (M CO 
00 C> r-i CO 00 O <M ^ 



«D CO ;o o 00 00 oo 00 

M CO ■«*< »0 (N CO rJH uo 



cu <3 



O 
W 
Pi 



00f0COO5COt> 



005f-I-^CD05 



lO I> 03 t-l CO CO 



CO»Ol^OOO(N 



^COrt*iOiO<D 



Tt^OlO-^(N^ 



CO oo'oo 00 00 00 

Tj* ■^ ITS lO CD CO 



144 



STEAM ENGINEERING 



u 

a 

is 

i- 
1§ 

J- 

c 

5s 

4) 

K- 

o 
o 


o 

§ 

a 

CO 


S888S88888S88S 


t^rtiooooiO'^cot^oototocoeo 

CO^iO<Ot^OOO(N»OTtt<©C^OOfO 


888S^g§Sg88888 


(N 00 O 00 Tjt (M CT> (35 00 O M OS N- O 
COfO»0»0«3t*000(N(N-*OiOO 


8§888888888§88 


00C<>(NOk0<MCDCCOC0NC0'«**O 
(N CO ■<*< lO »0 CO t^ O .-1 (N 0> « t^ 


23.00 
27.00 
34.00 
41.60 
45.60 
61.70 
63.40 
77.60 
91.60 
85.00 
101.00 
77.10 
111.00 
140.00 


8§S8g888§88288 


00 ^ 00 CO CO o o (M CO r- o »-i 00 o 

.-lN(MCOfO'*»OCOt>COOOCOOO'-l 


8g88g§8g8S8^88 


coioo-*coaicoTi<cooioO'<*-^o 

.-lr-<C^M<M(MCOrt<U5-*iO'«J<CO00 


§ 


(N 


Si?88g8ggg88gSi2 


00 0>C^^|O00j^gjHC.tr^O00^ 


1 

a 

6 


346 

411 

533 

612 

684 

776 

952 

1,164 

1,368 

1,276 

1,531 

1,170 

1,700 

2,117 




CO 


OOC5<N(M<M(M(N-<14Tt<COCDCO 

xli'x'x'x'x'x'x'x'x^ xl< X 

t^t^000000050-t<N--iCS»ON-* 



BOILER FEEDERS 



145 



HEAD NECESSARY TO GIVE REQUIRED 
PRESSURE 

When investigating this matter the 
engineer in charge of a plant may wish 
to know the head in feet, necessary to 
secure a given pressure, and while it is 
an "easy matter to calculate it by multi- 
plying the required pressure by 2.3, as 
this gives results that are practically 
correct, it saves time and labor to ob- 
tain these from a table, therefore the 
next table contains the "Feet Head"" or 
the height of a columnf of water necessary 
to produce the pressures given in the 
first column. 



PRESSURE AND HEAD OF WATER 



Pounds 


Feet 


Pounds 


Feet 


per square 
inch 


head 


per square 
inch 


head 


1 


2.31 


110 


253.98 


2 


4.62 


120 


277.07 


3 


6.93 


125 


288.62 


4 


9.24 


130 


300.16 


5 


11.54 


140 


323 . 25 


6 


13.85 


150 


346.34 


7 


16.16 


160 


369.43 


8 


18.47 


170 


392.52 


9 


20.78 


180 


415.61 


10 


23.09 


190 


438.90 


15 


34.63 


200 


461.78 


20 


46.18 


225 


519.51 


25 


57.72 


250 


577.24 


30 


69.27 


275 


643.03 


40 


92.36 


300 


692.69 


50 


115.45 


325 


750.41 


60 


138.54 


350 


808.13 


70 


161.63 


375 


865.89 


80 


184.72 


400 


922.58 


90 


207.81 


500 


1.154.48 


100 


230.90 


1.000 


2.308.00 



146 



STEAM ENGINEERING 



PRESSURE SECURED BY A GIVEN HEAD 

For other purposes it may be conven- 
ient to know how much pressure will be 
secured with a given head in feet, and 
while it can be calculated by multiply- 
ing the head by .434 as this will give 
practically correct results, as before 
mentioned, it is frequently much more 
convenient to consult a table and securer 
the required information at a glance. 

Working engineers and others some- 
times find it difficult to learn the use of 
tables, but a few hours spent in studying 
the subject will usually be sufficient to 
learn the value of these tables, especially 
if the more simple ones are studied first* 

HEAD AND PRESSURE OF WATER 



Head 


Pounds 


Head 


Pounds 


in feet 


per square 
inch 


in feet 


per square 
inch 


1 


.43 


110 


47.64 


2 


.87 


120 


51.97 


3 


1.30 


130 


56.30 


4 


1.73 


140 


60.63 


5 


2.17 


150 


64.96 


6 


2.60 


160 


69.29 


7 


3.G3 


170 


73.63 


8 


3.40 


180 


77.96 


9 


3.90 


190 


82.29 


10 


4.34 


200 


86.62 


20 


8.67 


225 


97.45 


30 


12.99 


250 


108.27 


40 


17.32 


275 


119.10 


50 


21.65 


300 


129.93 


60 


25.99 


325 


140.75 


70 


30.32 


350 


151.58 


80 


34.65 


400 


173.24 


90 


38.98 


500 


216.55 


100 


43.33 


1,000 


^33.30 



BOILER FEEDERS 147 

HORSE POWER REQUIRED TO RAISE 

WATER UNDER PERFECT 

CONDITIONS 

Actual conditions under which water 
is raised to different heights are always 
imperfect, as water passing through 
rough pipes, crooked passages in valves, 
around corners and past other obstruc- 
tions causes friction, which in turn 
wastes power and increases the cost of 
operation. 

The power thus developed, and used 
for no profitable purpose is a variable 
quantity, therefore when a plant is de- 
signed, this lost power can only be esti- 
mated after taking various conditions 
into consideration. After the plant is 
put into operation the amount of power 
wasted can be accurately determined 
by ascertaining the power actually used 
in raising the water, then subtracting 
the power that would be required if the 
conditions were perfect. 

When assuming perfect conditions all 
friction is neglected, and the power 
required is then determined by multiply- 
ing the weight of water raised per minute 
by the height to which it is elevated, 
and dividing the product by 33,000. 

While the horse power lost by friction 
is determined by subtracting the 
amount required under perfect condi- 
tions from the actual power used in 



148 STEAM ENGINEERING 

practice, this process does not deter- 
mine the friction of the pump as it in- 
cludes all lost power previously men- 
tioned, therefore it ought not to be 
charged to the pump only. 

When calculating power for perfect 
conditions, it is customary to use the 
term "theoretical horse power." While 
this is brief and convenient it is not cor- 
rect, therefore its use should be dis- 
couraged on all occasions. Correct 
theory always agrees with practice along 
these lines, for when all conditions are 
definitely known, theory takes them all 
into account. When they are not all 
known it is customary to ignore the un- 
certain points and call what is left the 
theoretical result. While custom "has 
established this precedent, it gives a 
wrong idea concerning the value of theo- 
retical calculations. 

The next table gives the power that 
would be required to raise water to va- 
rious heights, provided there was no 
friction to cause waste of power, or under 
perfect conditions. It is based on the 
following rule. 

Multiple the number of gallons 
pumped per minute, by the total height 
to which it is raised and divide by 4,000. 

The results secured by this rule and the 
one preceding it are practically identical, 
hence either may be used at pleasure. 



BOILER FEEDERS 



149 



POWER REQUIRED TO RAISE WATER 
Friction not included 



:Feet 


5 


10 


15 


20 


25 


30 


35 


Gal- 


1 












lons 


1 












5 


.006 


.012 


.019 


.025 


.0.31 


.037 


.044 


10 


.012 


.025 


.037 


.0.50 


.062 


.075 


.087 


15 


.019 


.037 


.056 


.075 


.094 


.112 


.131 


20 


.025 


.050 


.075 


.100 


.125 


.150 


.175 


25 


.031 


.062 


.093 


.125 


.156 


.187 


.219 


30 


.037 


.075 


.112 


.150 


.187 


.225 


.262 


35 


.043 


.087 


.131 


.175 


.219 


.262 


.306 


40 


,050 


.100 


.150 


, 200 


.250 


. 300 


.350 


45 


.056 


.112 


.168 


. 225 


.281 


. 337 


.394 


50 


.062 


.125 


.187 


.250 


.312 


.375 


.437 


60 


.075 


.150 


225 


. 300 


. 375 


.450 


.525 


75 


.093 


• .187 


.'^81 


, 375 


.469 


.562 


.656 


90 


.112 


.225 


.337 


.450 


.562 


.675 


.787 


100 


125 


.250 


.375 


.500 


.625 


.750 


.875 


125 


.156 


.312 


469 


. 625 


,781 


.9.37 


1.094 


150 


.187 


.375 


.562 


.750 


.937 


1.125 


1.312 


175 


.219 


.437 


.656 


.875 


1.093 


1.312 


1.531 


200 


.250 


.500 


.750 


1.000 


1.250 


1.500,1.750 


250 


.312 


.625 


937 


1 250 


1.562 


1.87512.187. 


300 


.375 


750 


1.125 


1.500 


1.875 


2.2.50 2.625 


350 


.437 


.875 


1.312 


1.750 


2.187 


2.625i3.062 


400 


.500 


1 ,000 


1 500 


9. 000 


?. 500 


3.00013.500 


500 


.625 


1.250 


1.875 


2.500 


3.125 


3.750 


4.375 



Feet 


40 


45 


50 60 j 75 1 90 


100 1 125 


Gal- 








1 1 






lons 








1 1 






5 


.05 


.06 


.06 


.07 


1 
.09 .11 


.12 


.16 


10 


,10 


.11 


,12 


.15 


,19 


,22 


.25 


.31 


15 


.15 


.17 


.1.9 


.22 


.28 


.34 


.37 


.47 


20 


.20 


.22 


.25 


.30 


.37 


.45 


.50 


.62 


25 


.25 


.28 


.31 


.37 


.47 


.56 


.62 


.It 


30 


.30 


.34 


.37 


.45 


.56 


.67 


.75 


.94 


35 


.35 


.39 


.44 


.52 


.66 


.79 


.87 


1.08 


40 


.40 


.45 


.50 


.60 


.75 


.90 


l.OOil.25 


45 


,45 


.51 


.56 


,67 


.84 


1.01 


1.12 


1.41 


60 


.50 


.56 


.62 


.75 


.94 


1.12 


1.25 


1 .56 


60 


.60 


.67 


.75 


.90 


1.12 


1.35 1.50 


1.87 


75 


.75 


.84 


.94 


1.12 


1.40 


1.69 


1.87 


2.34 



150 



STEAM ENGINEERING 



Power required to raise water — Continued 
Friction not included 



F.. 


40 1 45 


50 60 


75 


90 


100 


125 


Gal- 


1 












Ions 


1 












'90 


.90 


1.01 


1.12 


1.35 


1.68 


2.02 


2.25 


2.81 


100 


1.00 


1.12 


1.25 


1.50 


1.87 


2.25 


2.50 


3.12 


125 


1.25 


1.41 


1.56 


1.87 


2.34 


2.81 


3.12 


3.91 


\50 


1.50 


1.69 


1.87 


2.25 


2.81 


3.37 


3.75 


4.69 


175 


1.75 


1.97 


2.19 


2.62 


3.28 


3.94 


4.37 


^.47 


200 


2.00 


2.25 


2.50 


3.00 


3.75 


4.50 


5.00 


6.25 


250 


2.50 


2.81 


3.12 


3.75 


4.69 


6.62 


6.25 


7.81 


300 


8.00 


3.37 


3.75 


4.50 


5.62 


6.75 


7.50 


9.37 


350 


3.50 


3.94 


4.37 


5.25 


6.66 


7.87 


8.75 


10.94 


400 


4.00 


4.50 15.00 


6.00 


7.60 


9.00 


10.00 


12.60 


500 


5.00 


5.6216.25 


7.50 


&.37 


H.25 


12.60 


16.62 



Feet 


160 1 175 


200 


250 


300 1 350 


400 


Gal- 


1 






1 




lons 


1 










5 


.19 


.22 


.25 


.31 


.37 


.44 


.60 


10 


.37 


.44 


.50 


.62 


.75 


.87 


].00 


15 


.56 


.66 


.75 


.94 


1.12 


1.31 


1.60 


20 


.75 


.87 


1.00 


1.25 


1.50 


1.75 


2.00 


25 


.94 


1.09 


1.25 


1.56 


1.87 


2.19 


2.60 


30 


1.12 


1.31 


1.50 


1.87 


2.25 


2.62 


3.00 


35 


1.31 1.53 


1.75 


2.19 


2.62 


3.06 


3.60 


40 


1.50 


1.75 


2.00 


2.50 


3.00 


3.50 


4.00 


45 


1.69 


1,97 


2.25 


2.81 


3.37 


3.94 


4.60 


60 


1.87 


2.19 


2.50 


3.12 


3.75 


4.37 


5.00 


60 


2.26 


2.62 


3.00 


3.75 


4,50 


5.25 


6.00 


75 


2.81 


3.28 


3.75 


4.69 


6.62 


6.56 


7.60 


90 


3.37 


3.94 


4.50 


5.62 


6.75 


7.87 


9.00 


100 


3.75 


4.37 


5.00 


6.25 


7.50 


8.75 


10.00 


125 


4.60 


5.47 


6.25 


7,81 


9.37 


10.94 


12.60 


150 


5.62 


6.56 


7,50 


9,37 


11.25 


13.12 


15.00 


175 


6.56 


7.66 


8,75 


10.94 


13.12 


15.31 


17.50 


200 


7.50 


8.75 


10.00 


12.6C 


15.00 


17.50,20.00 


250 


9.37 


10.94 


12.60,15.72 


18.75,21.87 


25.00 


300 


11.25 


13.12 


15.00'l8.75 


22. 50126. 25 


30.00 


350 


13.12,15.31 


17.50]21.87 


26.25 


30.62 


35.00 


400 


15.00ll7.50 


20.00125.00 


30.00 


35.00 


40.00 


500 


18.75 


21.87 


25.00 


31.25 


37.60 


43.75 


60.00 



^ 



BOILER FEEDERS 151 

CONTENTS OF WATER PIPES 

When making calculations relating 
to boiler feeders, and for other purposes, 
it is frequently convenient to know the 
contents of water pipes used. In some 
cases it is necessary to suspend these 
pipes from joists or to place them on 
trusses, hence the weight to be carried 
should be known and provisions made 
for safely supporting it. In the case of 
very long pipes the water required to fill 
them, and the amount available when 
they are emptied, make more or less dif- 
ference in the results secured hence there 
ought to be a more convenient way at 
hand to determine this, than to calculate 
the area of pipe and its resulting capacity. 

If a vertical pipe is to be carried to an 
elevated tank, its weight when filled 
with water becomes a factor to be reck- 
oned with that should not be overlooked. 
The pressure due to the head of water 
in all such cases can be ascertained by 
consulting the table already given for 
this purpose. 

The following table contains data 
relating to pipes, that are valuable in 
this connection. The several columns 
will be readily imderstood from the ex- 
planation given herewith. 

N = The nominal inside diameter of 
pipe, or in other words the 
size as known to the trade. 



152 



STEAM ENGINEERING 



A=The actual inside diameter in 

inches. 
G=The number of U. S. gallons for 

each foot in length. 
W = The weight in pounds of water 

at standard temperature, or 

62 degrees Fah. for each foot 

in length. 
P = The weight of pipe perioot. 
C = The combined weight of water 

and pipe per foot. 

DIAMETER AND WEIGHT OF WATER 
PIPES 



N 


A 


G 


W 


P 


C 


H 


.270 


.0029 


.024 


,24 


.264 


1^ 


.364 


.0054 


.045 


.42 


.465 


/4 


.493 


.0099 


.083 


.56 


.643 


^2 


.622 


.0158 


.132 


.84 


.972 


Vi 


.824 


.0277 


.231 


1.12 


1.351 


1 


1.048 


.0447 


.373 


1.67 


2.043 


1J€ 


1.380 


.0777 


.648 


2.24 


2.888 


IJ^ 


1.610 


.1058 


.882 


2.68 


3.562 


2 


2.067 


.1743 


1.453 


3.61 


5.063 


2V2 


2.468 


.2483 


2.070 


5.74 


7.810 


3 


3.067 


.3835 


3.197 


7.54 


10.737 


ZV2 


3.548 


.5136 


4.291 


9.00 


13.291 


4 


4.026 


.6613 


5.512 


10.66 


16.172 


41/2 


4.508 


.8290 


6.910 


12.34 


19.250 


5 


5.045 


1.0380 


8.652 


14.50 


23.152 


6 


6.065 


1.5000 


12.503 


18.76 


31.263 


7 


7.023 


2.0120 


16.771 


23.27 


40.041 


8 


7.981 


2 . 5990 


21.6&4 


28.18 


49.844 


9 


8.937 


3.2590 


27.166 


33.70 


60.866 


10 


10.018 


4.0950 


34 . 134 


40.06 


74.194 


11 


11.000 


4.9370 


41.153 


45.02 


86.173 


12 


12.000 


5.8750 


48.972 


49.00 


97.972 


13 


13.250 


7.1630 


59.708 


54.00 


113.708 


14 


14.250 


8. 2850 


69.060 


58.00 


127.060 


15 


15.250 


9.4890 


79.097 


6^.00 


141.097 



BOILER FEEDERS 153 

DIRECT ACTING STEAM PUMPS 

The mechanical efficiency of a direct 
acting steam pump is low because the 
friction of it is excessive under many con- 
iditions found in practice, and although 
this may be reduced by expert manage- 
ment, it is not practicable to secure best 
results in a majority of cases. If more 
attention should be paid to the proper 
lubrication of these pumps it would 
prove to be a paying investment. The 
common brass lubricator (so called by 
courtesy) is still found in numerous 
cases, and it is filled at irregular intervals 
by the engineer or the fireman. One 
reason for this is that there is no way 
whereby they can tell at a glance 
whether it is full or empty, and it is 
neglected as a natural consequence. 

Another bad feature of the appliance 
is that after it is filled with oil, there is 
no principle involved in its operation 
whereby it feeds the oil regularly, con- 
sequently it frequently becomes empty 
'i in a few minutes, after which the valves 
il .and the steam pistons must run without 
i||oil until more happens to be supplied. 
3i| At other times the oil remains in the 
1 lubricator for a long time, leaving the 
■ internal moving parts to run dry when 
c they need lubrication. Every direct 
- : acting steam pump should be fitted with 



i 



154 STEAM ENGINEERING 

a sight feed lubricator that feeds oil as 
regularly as it is fed to the main engine. 

The thermal efficiency of the direct 
acting steam pump is low, because but 
little of the steam supplied to it is actu- 
ally used to produce power, and some 
of it is condensed in the cylinders owing 
to the slow piston speed which condi- 
tions at the water end impose, and 
nearly all of the remainder passes out 
through the exhaust pipe. 

Notwithstanding these unfavorable 
conditions this boiler feeder is used in 
thousands of plants, and when all points 
are considered it is one of the best ma- 
chines in use at the present time for this 
purpose. 

The latter statement apparently con- 
tradicts the two which precede it, but 
the reasons for it are that this type of 
pump can easily be regulated to supply 
any quantity of water wanted from 
nothing to the full capacity of the ma- 
chine, aiid the exhaust steam from it is 
frequently used to good advantage 
elsewhere. 

This secures economy at the boiler by 
maintaining a constant water level, 
and where live steam must be used for 
heating provided there is not enough 
exhaust steam to answer the purpose, 
low thermal efficiency is not detrimental. 



BOILERS FEEDERS 



155 



SINGLE PUMPS 



Fig. 20 illustrates a single direct act- 
ing steam pump. This is an ideal kind, 




provided means are supplied for revers- 
ing the piston when it has reached the 



156 STEAM ENGINEERING 

end of a stroke. As there is no fly- 
wheel used, nor any substitute for it, 
there is no momentum developed that 
can be used to operate a steam valve 
and cause it to admit steam to the other 
end of the cylinder, therefore, a device 
must be employed that will operate a 
secondary valve and admit steam to a 
supplemental piston, and when this 
moves it takes the main steam valve 
with it, thus securing the desired effect. 
A great variety of devices have been 
patented and used for this purpose with 
more or less success until perfect ma- 
chines are now manufactured, leaving 
nothing to be desired along this line. 
Many working engineers remember the 
time when a perfect working single 
pump was the exception rather than the 
rule, but competition has changed these 
conditions until satisfactory single 
pumps are the rule to which there are 
very few exceptions. 

It is claimed that some of this type of 
pumps are so well designed and con- 
structed that they have but ^8 inch 
clearance at each end of the stroke, 
still the steam piston does not strike the 
cylinder heads, neither does it make 
short strokes under ordinary working 
conditions. 

A single pump must of necessity cease 
drawing and forcing water at the end of 



BOILER FEEDERS 157 

each stroke, unless there is a high Hft by 
suction, or the speed is excessive, caus- 
ing a body of water in the pipes to attain 
more or less momentum, so that when 
the water piston stops at the end of each 
stroke, the body of water continues its 
motion at reduced speed, until another 
stroke is begun which again forces the 
water forward. This action does not 
always prevent a check valve in the dis- 
charge pipe from seating frequently, 
and causing more or less shock and jar 
during the operation, but this is not 
objectionable at slow speed. Pump 
makers do not recommend such high 
speeds now as formerly, which is a move 
in the right direction. When the capac- 
ity of a pump is given, the number of 
strokes necessary to secure the given 
amount of water should be noted, es- 
pecially if two or more pumps are to be 
compared, as otherwise the comparison 
may not be fair to both parties. 

Where competition is great, there is a 
tendency to rate a pump at an exces- 
sively high speed, thus giving the manu- 
facturer who adopts these tactics a tem- 
porary and unfair advantage over a 
more conservative competitor, but care- 
ful comparison of all sizes will do much 
towards correcting this evil. 

The next table gives sizes of pumps 
that are recommended for boiler feeding 



158 STEAM ENGINEERING 

and pumping water for other purposes 
under heavy pressure. 

The comparative sizes of steam and 
water pistons, also the length of strokes 
will be satisfactory in practice. When 
selecting a pump for a certain boiler it 
should be remembered that a pump 
must supply water for the actual service 
required of this boiler, and not for an 
arbitrary rating which may or may not 
be correct. 

It is useless to compare sizes of pipes 
for different pumps, as a pump may not 
take an excessive amount of steam to 
run it, because provision is made for 
fitting it with a large steam pipe, neither 
will a pump draw a large quantity of 
water solely on account of having a 
large suction pipe. 

Letters at the head of these columns 
refer to sizes and capacities as follows: 

S = Diameter of steam piston 
W= " " water « 

L = Length of stroke. 
G = Gallons per stroke. 
N = Number of strokes per minute 

at which the pump can be 

operated. 
C = Capacity at given speed. 



BOILER FEEDERS 159 

SIZE AND CAPACITY OF SINGLE PUMPS 



s 


W 


L 


G 


N 


c 


3 


IH 


3 


.031 


80 to 175 


2.48 to 5.42 


3 


2 


3 


.04 


80 


' 175 


32 " 7 


4 


2H 


5 


.106 


75 


' 150 


8 " 16 


5 


3 


6 


.183 


75- 


' 150 


14 " 27 


5y2 


33^ 


7 


.291 


75 


' 150 


22 " 44 


6M 


4 


8 


.44 


75 


' 150 


36 " 66 


7H 




9 


.55 


75 


* 125 


41 " 69 


7H 


43^ 


9 


. .62 


75 


' 125 


46 " 78 


8H 


5 


10 


.85 


75 


' 125 


64 " 106 


8.4 


5H 


10 


1.02 


75 


' 125 


75 !' 128 


10 


6 


12 


1.46 


75 


' 125 


110 " 183 


10 


QH 


12 


1.72 


75' 


125 


129 " 215 


12 


7 


12 


2.00 


75 


' 125 


150 " 250 


12 


7H 


12 


2.29 


75 


' 125 


172 " 286 


12 


7 


16 


2.66 


50 


' 100 


133 '- 266 


12 


•7H 


16 


3.06 


50 


' 100 


153 " 306 


14 


7J^ 


12 


2.29 


75 


' 125 


172 " 286 


14 


8 


12 


2.61 


75 


' 125 


196 " 326 


14 


8K 


12 


2.94 


75 


' 125 


221 " 367 


14 


8 


16 


3.48 


50 


' 100 


174 •' 348 


14 


8H 


1& 


3.93 


50 


' 100 


197 " 394 


14 


9 


16 


4.40 


50 


' 100 


220 " 440 


16 


9 


16 


4.40 


50 


' 100 


220 " 440 


16 


10 


16 


5.44 


50 


' 100 


272 " 544 


18 


12 


16 


7.83 


50 


' 100 


392 " 784 



REMARKS ON SINGLE PUMPS 

The manufacturers of these pumps 
state that they can be run at "a much 
higher speed than their catalogues state, 
in case of emergency, but are careful to 
add that for continuous work against 
high pressure, it is better to run them 
much slower, all of which is herewith 
recommended for intelligent considera- 
tion of all who are interested in giving 
pumps a reasonable chance for long life 
and much usefulness. 



160 STEAM ENGINEERING 

A pump that is designed to run with- 
out outside valve gear can usually be 
run at a high speed without excessive 
vibration, but that may not be a good 
reason for careless use of it in running 
it at excessive speed for a time and then 
letting it stand for perhaps an equal 
time when it could be run at a slow 
speed continuously, thus greatly pro- 
longing its life. 

On the other hand a single pump that 
is fitted with a valve gear in which a blow 
is struck every time that the piston 
is reversed, should never be run at a 
high speed. 

The pump illustrated in Fig. 20 is a 
submerged piston pump, because both 
suction and discharge valves are above 
the water piston, hence while in opera- 
tion it must always be covered by water. 

The natural result of this is that it has 
superior lifting power, as the water seal 
so provided assists in keeping it air 
tight. On this account a pump of this 
style is recommended where it is neces- 
sary to lift water 20 feet or more. 

LIFTING POWER OF PUMPS 

A pump cannot lift water higher than 
a column of water will stand to balance 
the weight of the atmosphere at that 
particular locality. This is about 14.7 
pounds at sea level, which is equal to a 



BOILER FEEDERS 



161 



column of water 33.95 feet high. The 
practical working lift is much less, be- 
cause it is impossible to create and main- 
tain a perfect vacuum in the suction pipe, 
also owing to the fact that the pump 
,must be set lower in order to allow the 
water to run into it when in operation, 
after it has been lifted. 

Furthermore, the limit of suction 
power of a pump is not the same in all 
localities, because the weight of the 
atmosphere becomes less as high ground 
is reached. The next table contains 
information on this point. 



05 CO 
^ g 



JU 

H 






1°§ 
111 


25 ft. 
24 ft. 
23 ft. 
21ft. 
20 ft. 
19 ft. 
18 ft. 
17 ft. 


1- 
•III 


33.95 ft. 
32.38 ft. 
30.79 ft. 
29.24 ft. 
27.76 ft. 
26.38 ft. 
25.13 ft. 
22.82 ft. 


Is 


14.70 lbs. 
14.02 lbs. 
13.33 lbs. 
12.66 lbs. 
12.02 lbs. 
11.42 lbs. 
10.88'lbs. 
9.88 lbs. 


1 


Sea level 
J^ mile or 1,320 ft. 
H mile or 2,640 ft. 
5^ mile or 3,960 ft. 

1 mile or 5,280 ft. 
IM mile or 6,600 ft. 
IJ^ mile or 7,920 ft. 

2 miles or 10,560 ft. 



162 



STEAM ENGINEERING 



For the foregoing table the water is 
assumed to be at standard temperature 
or 62 degrees Fah., and although it may 
vary several degrees in either direction 
without affecting the final result, the 
temperature can easily be raised high 
enough, to change the limit of suction 
power because reducing atmospheric 
pressure on the surface of a body of 
water lowers its boiling point, hence it 
will evaporate, or in other words allow 
the pipe to fill with vapor instead of 
raising a solid body of the hot water. 

The next table gives the maximum 
depth of suction power, and as the prac- 
tical limit for general service is much less 
than the maximum, it follows that very 
hot water must run to a pump by 
gravity or else it will not be delivered 
to the boiler. 



THE LIMIT OF SUCTION POWER 




For Hot Water 




Temper- 
ature 


Absolute 


Vacuum 


Limit of 
suction 


Fah. 


pressure 


in inc hes 


power 


101.4 


1 


27.88 


31.6 


126.2 


2 


25.85 


29.3 


144.7 


3 


23.81 


27.0 


153.3 


4 


21.77 


24.7 


162.5 


5 


19.74 


22.4 


170.3 


6 


17.70 


20.1 


177.0 


7 


15.66 


17.8 


183.0 


8 


13.63 


15.5 


18S.4 


9 


11.59 


13.2 


193.2 


10 


9.55 


10.9 


197.6 


11 


7.51 


8.5 


201.9 


12 


5.48 


6.2 


205.8 


13 


3.44 


3.9 


209.6 


14 


1.40 


1.6 



BOILER FEEDERS 163 

DUPLEX PUMPS 

The duplex direct acting steam pump 
owes its existence and great popularity 
largely to the fact that before it was 
brought into extensive use, the single 
pumps then in service were considered 
unreliable, as they frequently failed to 
operate when most needed. 

This does not refer to any particular 
kind, neither does it necessarily mean 
that there were no good ones in evidence 
at that time. Although this style of 
pump has several defects that have not 
been eliminated, it will always go when 
steam is turned on, provided it has not 
been purposely disabled, or is badly out 
of repair. 

Its mechanical efficiency is low and its 
thermal efficiency is not high, yet it is 
popular and probably always will be. 
If it does discharge an enormous quan- 
tity of steam for the power it develops, 
in many cases this steam can still be used 
for heating buildings in winter, and 
dry kilns and other departments in sum- 
mer, thus reducing the real waste to a 
low percentage. 

Fig. 21 is a side view in section of one 
of these pumps, fitted with submerged 
pistons. The whole machine is' very 
simple, hence not liable to be deranged 
in rough service. 



164 STEAM ENGINEERING 




As the crosshead of one pump oper- 
ates the valve of the other, it is impos- 
ible to run one of them separately in 



BOILER FEEDERS 165 

case the other is disabled from any 
cause, for a duplex pump is only two 
single pumps located on one base, but 
they are not fitted with independent 
valve gears. 

SETTING THE VALVES OF DUPLEX PUMPS 

This process seems to be a great mys- 
tery to those who have not given the 
matter due attention, but when the 
principles involved are understood, it 
becomes an easy matter to apply them 
in practice. 

By observing Fig. 21 carefully it will 
be noted that the steam piston is in the 
center of its travel, hence the first move 
to make in setting the valves is to move 
the steam piston as far in one direction 
as it will go, then mark the rod with a 
lead pencil at the edge of the stuffing 
box, or at some other suitable place. 
. Next move it as far as it will go in the 
opposite direction, and mark it in the 
same way. Place another mark cen- 
tral between these two and when the 
latter is flush with the face of the stuffing 
box gland, the piston is in the center of 
its travel. 

Remove the steam chest cover and 
place the flat slide valve in the center 
of its travel, and adjust the lost miotion 
at the valve rod so that it will be equally 
divided on each side of the valve. 



166 STEAM ENGINEERING 

The same device is not used for this 
purpose on all pumps, hence only gen- ; 
eral directions can be given at this time. 
If the device is adjustable, allow enough 
lost motion on each side of the valve to 
equal one-half of the width of the steam 
port. . Repeat the operation for the 
other pump and the valves are set. If 
the stroke proves to be too long, allow- 
ing the piston to strike the cylinder 
head, reduce the lost motion on the 
valve rod. If the stroke is too short, 
give more lost motion. 

BRASS AND FIBROUS PACKED PISTONS 

Fig. 21 illustrates a piston that is 
made in several adjustable pieces and 
packed with a suitable number of rings 
of square fibrous packing. One advan- 
tage of this design is that if pieces of iron 
scale or other foreign matter from the 
inside of heating pipes and radiators 
finds its way to the cylinder of such a 
pump when it is taking the returns from 
heating systems, it is not liable to score 
the cylinder, although there is some 
danger of it. 

Another advantage is found in the 
fact that when this packing is worn out, 
it can easily be replaced by the engineer, 
without making it necessary to employ 
a machinist to make repairs, or send 



BOILER FEEDERS 167 

some of the parts back to the manufac- 
turer. 

One disadvantage of this design is 
that if an inferior grade of packing is 
used by mistake or otherwise, it will not 
last long, and it may fail when needed 
badly, although it usually gives warning 
to the observing engineer, by partially 
failing before it wholly disappears. 

Another disadvantage is that it is 
impossible to always pack the two pis- 
tons of a duplex pimip exactly alike, 
therefore if both of them are too tight, 
they will both make short strokes, and 
greatly reduce the capacity of the pump 
for a given number of strokes per min- 
ute. If one is packed tighter than the 
other, then one will give a shorter stroke 
than it was intended for, reducing the 
capacity and making an unfavorable 
appearance. 

Fig. 22 illustrates a pump in which 
no fibrous packing is used for the water 
end. It is fitted with a brass plunger 
made in one piece, working in a brass or 
composition ring. 

The advantages of this design are that 
it proves durable in practice, as only 
comparatively hard metal is used in its 
construction, and as it is not adjustable 
there is no danger of packing one, or 
both of them on a duplex pump, too 



STEAM ENGINEERING 




BOILER FEEDERS 169 

tight, thus increasing friction and prov- 
ing unsatisfactory. 

Its disadvantages are that when it 
does need repairs they cannot be made 
by the average working engineer with 
the tools commonly found at hand, 
therefore it will usually be used in poor 
condition for some time before it is con- 
sidered actually in need of repairs, as 
new parts must be sent for and put in 
place. If such a pump fails at a critical 
time it cannot be repaired quickly, un- 
less the extra parts are kept in stock for 
such an emergency. 

The piston in Fig. 22 is not submerged, 
hence this kind of a pump is not suitable 
for a high suction lift. Water passes 
through it with the least possible fric- 
tion, as it enters the suction chamber, 
passes directly up through the lower 
valves and is discharged through the 
upper valves into the discharge chamber 
without changing the direction in which 
it travels. 

AH of the pumps illustrated are suita- 
ble for either hot or cold water, but 
valves that are appropriate for cold 
water will soon fail if hot water is ad- 
mitted to them. Hot water valves 
will answer for cold water, provided 
they fit their seats perfectly. As hot 
■ water valves are usually made of com- 
paratively hard material, a. slightly un- 



170 STEAM ENGINEERING 

even surface on either the valve or its 
seat will cause a leak and prevent the 
pump from working properly. 

The next table contains the size and 
capacity of duplex pumps. The several 
columns are designated by the same 
letters used for the table of size and 
capacity of single pumps, and they refer 
to the same parts and conditions except 
the last. This is added in order to show 
the diameter of water cylinder that 
must be used to secure equal capacity 
in a single pump, assuming that both 
pistons of the duplex pump make full 
strokes. 

S = Diameter of steam piston. 

W = " " water " or 

plunger. 

G = Gallons of water per single 
stroke. 

N = Number of strokes per minute, 
at which each piston can be 
operated. 

C = Capacity of both cylinders at 
given speed. 

H = Diameter of the water piston of 
a single pump that will equal 
both water pistons of a du- 
plex pump of given size. 



BOfLER FEEDERS 171 

SIZE AND CAPACITY OF DUPLEX PUMPS 



s 


W 


L 


G 


N 


C 


H 


.2 


Wi 


2M 


.013 


100 to 300 


2 to 7 


IVz 


3 


2 


3 


.04 


100 ' 


250 


8 


' 20 


2Vi 


^V2 


2K 


4 


.10 


100 * 


200 


20 


' 40 


4 


^K 


3H 


5 


.20 


100 ' 


200 


40 


' 80 


5 


6 


4 


6 


.33 


100 " 


150 


70 


' 100 


5H 


iy2 


4>^ 


6 


.42 


100 ' 


150 


85 


' 125 


6M 


ly, 


5 


6 


.51 


100 ' 


150 


100 


' 150 


7 


7H 


41^ 


10 


.69 


75 ' 


125 


100 


' 170 


^n 


p 


5M 


10 


.93 


75 * 


125 


135 


' 230 


734 


10 


6 


10 


1.22 


75 ' 


125 


180 


' 300 


834 


10 


7 


10 


1.66 


75 ' 


125 


245 


' 410 


9Ji 


12 


7 


10 


1.66 


75 ' 


125 


245 


' 410 


9J^ 


14 


7 


10 


1.66 


75 ' 


125 


245 


' 410 


9^8 


12 


8H 


10 


2.45 


75 ' 


125 


365 


' 610 


12 


14 


8H 


10. 


2.45 


75 ' 


125 


365 


' 610 


12 


16 


8^ 


10 


2.45 


75 ' 


125'365 


' 610 


12 


181^ 


8i^ 


10 


2.45 


75 • 


125 


365 


' 610 


12 


20 


8H 


10 


2.45 


75 ' 


125 


365 


' 610 


12 


12 


10 K 


10 


3.57 


75 * 


- 125 


530 


' 890 


14Ji 


14 


10^ 


10 


3.57 


75 ' 


125 


530 


' 890 


14^ 


16 


10 1^ 


10 


3.57 


75 * 


125 


530 


' 890 


14M 


18J^ 


ioj<r 


10 


3.57 


75 ' 


125 


530 


' 890 


14 M 


20 


lOM 


10 


3.57 


75 ' 


125 


530 


• 890 


14 >i 


14 


12 


10 


4.89 


75 * 


125 


730 


'1220 


17 


16 


12 


10 


4.89 


75 ' 


125 


730 


'1220 


17 



COMPQUNB' STEAM PUMPS 

Where fuel is expensive and the ex- 
haust steam from a pump cannot be 
used for other purposes, the compound 
pump, either single or duplex is recom- 
mended, as it will save about 30 per 
cent, of fuel over what a single pump 
requires to do the same work. 

The principle involved in the use of 
steam in these two kinds of pumps is 
that in. the single pump (which may be 



172 STEAM ENGINEERING 

either single or duplex), steam, is admit- 
ted for the whole length of every stroke, 
and then exhausted at practically the 
same pressure that it had at the begin- 
ning of the stroke, hence there is great 
waste of heat. 

In the compound pump (which may 
be either single or duplex), steam is ad- 
mitted to the high pressure cyHnder and 
after its work is completed here, it is 
exhausted into the low pressure cylin- 
der (which is much larger than the high 
pressure), where it is expanded to a 
comparatively low terminal pressure. 
As work is done during the expanding 
process the results so far as the consump- 
tion of fuel is concerned, are satisfac- 
tory. 

Fig. 23 illustrates this type of pump, 
and examination of it shows that it is 
not a complicated machine, as it can be 
cared for by anybody who is competent 
to take care of a single pump to do the 
same work. 

The use of steam in the second cylin- 
der causes back pressure in the first, 
and this l,e;ssens the economy that would 
be secured if this back pressure could be 
eliminated, but owing to the fact that it 
opposes the progress of a small piston 
and assists a much larger one, there is a 
material gain by the process. 



BOILER FEEDERS 173 




174 STEAM ENGINEERING 

The next table contains the dimen- 
sions and capacity of compound steam 
pumps to which the following explana- 
tion applies: 

H P = Diameter of high pressure 
cylinder. 

L P = Diameter of low pressure cyl- 
inder. 

W C = Diameter of water cylinder. 

S = Stroke in inches. 

G = Gallons per stroke. 

N = Number of strokes per minute 
for a single pump that can be 
secured in every day prac- 
tice without excessive wear. 
The number will be doubled 
in case of a duplex pump. 

C = Capacity in pounds per minute 
for a single pump, at given 
speed. 

D = Capacity in pounds per minute 
for a duplex pump giving 
full strokes, at given speed. 

Attention is called to the fact that the 
capacity of each is high because the pis- 
ton speed is high, although the number 
of strokes per minute is comparatively 
low. These desirable features are se- 
cured by adopting a long stroke. 



BOILER FEEDERS 



175 



SIZE AND CAPACITY OF qOMPOUND 
PUMPS 



HP 


LP 


WC 


S 


G 


N 


c 


D 


4 


7 


4 


10 


.55 


100 


456 


912 


5y2 


8 


4 


10 


.55 


100 


456 


912 


6 


10 


4 


10 


.55 


100 


. 456 


912 


7 


12 


4 


10 


.55 


100 


456 


912 


5H 


8 


5 


10 


.85 


100 


705 


1,410 


6 


10 


5 


10 


.85 


100 


705 


1,410 


7 . 


12 


5 


10 


.85 


100 


705 


1.410 


8 


12 


5 


10 


.85 


100 


705 


1.410 


6 


10 


6 


10 


1.22 


100 


1,012 


2,024 


7 


12 


6 


10 


1.22 


100 


1,012 


2.024 


8 


12 


6 


10 


1.22 


100 


1,012 


2,024 


9 


14. 


6 


12 


1.47 


100 


1,220 


2,440 


10 


16 


6 


12 


1.47 


100 


1,220 


2,440 


12 


18 


6 


12 


1.47 


100 


l,220i 2,440 


7 


12 


7 


10 


1.66 


100 


1,377 


2,754 


8 


12 




10 


1.66 


100 


1,377 


2.754 


9 


14 


7 


10 


1.66 


100 


1,377 


2.754 


10 


16 


7 


12 


2.00 


100 


1,660 


3,320 


12 


18 


7 


12 


2.00 


100 


1,660 


3,320 


14 


20 


7 


■12 


2.00 


100 


1.6601 3,620 


8 


12 


S 


12 


2.61 


100 


2,16614,332 


9 


14 


8 


12 


2.61 


100 


2,16614,332 


10 


16 


8 


12 


2.61 


100 


2,166 4,332 


12 


18 


8 


12 


2.61 


100 


2,166 4,332 


14 


20 


8 


12 


2.61 


100 


2,166 4,332 


9 


14 


9 


12 


3.30 


100 


2.739 


5.478 


10 


16 


9 


12 


3.30 


100 


2,739 


5,478 


12 


18 


9 


12 


3.30 


100 


2,739 


5,478 


14 


20. 


9 


12 


3.30 


100 


2,739 


5,478 


16 


24 


9 


12 


3.30 


100 


2,739 


5,478 


10 


16 


10 


12 


4.08 


100 


3,386 


6,772 


12 


IS 


10 


12 


4.08 


100 


3.386 


6,772 


14 


20 


10 


12 


4.08 


100 


3,386 


6,772 


16 


24 


10 


12 


4.08 


100 


3,386 


6,772 


10 


16 


10 


14 


4.75 


80 


3,154 


6.308 


12 


18 


10 


14 


4.75 


80 


3,154 


6,308 


14 


20 


10 


14 


4.75 


80 


3,154 


0,308 


16 


24 


10 


14 


4.75 


so 


3,154 


6,308 


10 


16 


10 


20 


6.80 


80 


4,515 


9,030 


12 


18 


10 


20 


6.80 


80 


4,515 


9,030 


14 ■ 


20 


10 


20 


6.80 


80 


4.515 


9,030 


16 


24 


10 


20 


6.80 


so 


4,515 


9,030 


12 


IS 


12 


14 


6.85 


80 


4,548 


9,096 


14 


20 


12 


14 


6.85 


80 


4,548 


9,096 


16 


24 


12 


.14 


6.85 


80 


4.548 


9,096 


12 


18 


12 


24 


11.75 


50 


4,872 


9,744 


14 


20 


12 


24 


11.75 


50 


4,872 


9.744 


16 


24 


12 


24 


11.75 


50 


4,872 


9.744 


18 


30 


12 


24 


11.75 


50 


4,8721 9.744 



176 STEAM ENGINEERING 

THE SPEED OF STEAM PUMPS 

The most difficult point to decide 
when ordering a steam pump, is the 
speed at which it will give the best re- 
sults when everything is considered. 

The purchaser does not want to buy a 
pump that is too large for the service 
required, as that means, extra expense, 
hence he is inclined to take the maker's 
statement of the speed at which it can 
be run in case of emergency, for what 
will give good results in practice; but this 
is a mistaken idea because excessive 
speed wears a pump more than it would 
be 'worn if it were run at a slow speed 
for a much longer time. 

Furthermore, it greatly increases the 
danger of failure on account 'of severe 
shocks and jars. 

It is claimed that a piston speed of 
100 feet per minute is the highest limit 
that can be allowed for a pump, and it 
is quite safe to say that it should never 
be exceeded for boiler feeding purposes, 
and that in many cases the limit ought 
to be placed at 50 feet. 

It is a self-evident truth, although it 
may not always be remembered, that a 
high piston speed never damages a pump 
while the piston is swiftly moving in one 
direction, but when its motion is quick- 
ly reversed the resulting concussion 



BOILER FEEDERS 



177 



causes rapid wear and loosening of the 
joints. 

The lesson to be leagied from these 
observations, and practical experience 
along the same line is that a pump 
should always be designed with a long 
stroke. It is not practicable to lay 
down a cast iron rule for the length of 
stroke, but it should be from 2 to 3 times 
the diameter of the water cylinder. 

Some pumps are designed with a 
stroke longer than this indicates, as it is 
4 or even 5 times the diameter of the 
water cylinder, but they are designated 
as a special kind. There is no good 
reason why the stroke of boiler feeders 
should not be made much longer than it 
now is in many cases. 

The following, dimensions are taken 
from the catalogue of a prominent firm 
who have long been engaged in the 
manufacture of pumps of many kinds. 
They would make first class boiler 
feeders. 

LONG STROKE PUMPS 



Diameter of 


Diameter of 


Length 


steam cylinder 


water cylinder 


of stroke 


10 


5 


25 


10 


6 


25 


12 


7 


25 


14 


9 


33 


16 


10 H 


33 


18 


12 


38 



-178 



STEAM ENGINEERING 



The next table contains the nuniber of 
strokes that pumps must make to attain 
speeds from 50 to 125 feet per minute, 
with strokes ranging from 3 to 18 inches 
in length. 

It not only gives valuable information 
along this line, but it enables the reader 
to make comparisons readily which show 
that a statement of the piston speed of 
a pump does not always give an intelli- 
gent idea of its operation. 

For illustration, a pump with an 18 
inch stroke will attain a piston speed of 
100 feet per minute by making 67 strokes, 
while another with a stroke of 5 inches 
must make 240 strokes to attain the same 
speed. A pump with a 4-inch stroke at- 
tains a speed of only 50 feet per minute 
when making 150 strokes, but when a 
pump with a 10-inch stroke makes 150 
strokes it has a pi^ston speed of 125 feet. 



NUMBER OF STROKES REQUIRED 


ro 


ATTAIN A GIVEN PISTON SPEED 


u 


Stroke in inches 


ll 


3 


4 


5 


6 1 7 1 8 


10 1 12 1 15 1 18 


c^E 


Number per minute 


50 


200 


150 


120 


100 


86i 75 


60 


50 


40 


33 


55 


220 


165 


132 


110 


94 


82.5 


66 


55 


44 


37 


60 


240 


180 


144 


120 


103 


90 


72 


60 


48 


40 


65 


260 


195 


156 


130 


111 


97.5 


78 


65 


52 


43 


70 


280 


210 


168 


140 


120 


105 


84 


70 


56 


47 


75 


300 


225 


180 


150 


128 


112.5 


90 


75 


60 


50 


80 


320 


240 


192 


160 


137 


120 


96 


80 


64 


53 


85 


340 


255 


204 


170 


146 


127.5 


102 


85 


68 


57 


90 


360 


270 


216 


180 


154 


135 


108 


90 


72 


60 


95 


380 


285 


228 


190 


163 


142.5 


114 


95 


76 


63 


100 


400 


300 


240 


200 


171 


150 


120 


100 


80 


67 


105 


420 


315 


252 


210 


180 


157.5 


126 


105 


84 


70 


110 


440 


330 


264 


220 


188 


165 


132 


110 


88 


73 


115 


460 


345 


276 


230 


197 


172.5 


138 


115 


92 


77 


120 


480 


360 


288 


240 


206 


180 


144 


120 


96 


80 


125 


500 


375 


300 


250 


214 


187.5 


150 


125 


100 


83 



BOILER FEEDERS 



179 



::^;^ :^:^;^ :i^:s;§^ :5t^:?^ :it-^;^ 



180 



STEAM ENGINEERING 







g^gg8^^2S^§^2§g§gS§ 




183 
215 
249 
286 
326 
368 
413 
460 
510 
562 
617 
734 
861 
999 
1.147 
1,305 
1.652 


B 

c 
1 

u 

O. 

c 
S 
.2 


o 
o 


oc(roa>»o>-it>.-^c^oooo«oiot-'0»oo> 


146 
172 
199 
229 
261 
294 
330 
368 
408 
449 
493 
587 
689 
799 
918 
1,044 
1.321 


§ 


S^S;i?8^^??§S3j:?t2Sf2§§i: 




§ 


OOtHOOWOOOOOCD-^OOOOOO 
iOOOiCDOiCOCC'OTj<C00500t^-«l<OiO 


117 
137 
159 
183 
208 
235 
264 
294 
326 
329 
394 
470 
551 
639 
734 
835 
1.057 


o 


T-tcD-*iooo^coioor>-t^'Oioo50»oro 
t^<oo50i>ir>:)cot^coooiO(M(£>r^to.-Hco 


C<10050iMCO'-it^>0-*iO>-i(MOC<Ji-liO 


o 

CO 


^^g^^^g^sss^^g^ss^s 


^sSSSSsiiiiiSSsig 


g 






o 


SgS:§!5S2?^gg^8§g§S§ 


kOOt^050^co'j<ot^05cor^'-icr>— i(M 




c 


\M \c-» \N s^ <-fv 

«DtOr^t^0000O5OSrH.-l.-i.-.r-.rHr-l.-..-( 



1^ 



BOILER FEEDERS 181 

EXPLANATION OF THE PRECEDING 
TABLES 

An explanation of the practical opera- 
tion of the two foregoing tables will 
assist the reader in making ready appli- 
cation of them and point out interesting 
features that otherwise might be over- 
looked. 

Suppose that in a certain case 160 
gallons of water are wanted per minute, 
and the proper size of pump is required. 
The piston speed is limited to 50 feet per 
minute for ordinary use, which is to be 
doubled in case of emergency. Taking 
the table of "Capacities of Pumps in 
Gallons," finding the column under 50 
feet per minute and following it down 
we find that the next quantity above 
160 is 165.24 which will answer the pur- 
pose, as the quantity delivered should 
never be less than the requirements. 
In the column headed "Diameter in 
inches," opposite the quantity delivered 
as above stated is a figure 9 indicating 
that the water piston must (or at least 
ought to) be 9 inches in diameter. 

Referring to the table entitled "Ntim- 
ber of Strokes Required to Attain a 
Given Piston Speed," we find that if a 
stroke of 15 inches is adopted it will only 
be necessary to make 40 strokes per 
minute^to attain a piston speed of 50 



182 STEAM ENGINEERING 

feet, thus securing conditions that result;: 
in ease of operation, and durability of ;i 
the machinery, as we will have a 9X15,! 
inch pump running 40 strokes per 
minute. 

Again suppose that 280 gallons of 
water are wanted per minute and it is 
to be secured with a pump that does not ; 
make more than 50 strokes per minute, 
and the piston speed is limited to 70 
feet. What is the proper size? 

Following down the column under 70 
until 285.60 gallons are noted we find 
that a piston 10 inches in diameter will 
give the desired quantity. In the other 
table we find 70 in the column entitled 
"Speed per minute," and opposite this 
the only number below 50 is 47 strokes 
per minute, and by following this col- 
umn up to the top we find that the 
proper stroke is 18 inches, therefore the 
pump should be 10X18 inches. 

It may be necessary to determine the 
capacity of a pump that is available for 
a certain place, provided it is large 
enough. Suppose that the water cylin- 
der of this pump is 4 inches in diameter 
with a stroke of 8 inches and it is con- 
sidered safe to operate it at 120 strokes 
per minute for maximum speed. What 
is the capacity of this pump? 

Under 8 in the proper column we find 
120 and by following the line to the left 



BOILER FEEDERS 183 

the piston speed is found to be 80 feet 
per minute, which is not excessive for a 
maximum speed. In the other table 
we find 80 and by following it down 
until it intersects the line where 4 is 
located at the left hand, the capacity is 
found to be 52.22 gallons per minute. 

A pump with a water cylinder 4 inches 
in diameter and a stroke of 4 inches, 
delivers 60 gallons of water per minute 
when running 300 strokes, but this 
speed is excessive, causing unecessary 
wear and much trouble. How fast 
will a 6 X 8 inch pump run if used to 

! supply the same quantity of water? 

I In the column entitled "Diameter in 
inches," we find 6 and following out that 
line until we find 73.44 gallons, which is 

i the next number above 60, Following 
this coliunn upward shows that in order 
to deliver this quantity of water the 
piston speed will be 50 feet per minute. 
P-ef erring to the other table we find that 
if a pump with a stroke of 8 inches is run 
50 feet per minute it will make 75 strokes 
giving satisfactory service, therefore if 
the 4X4 inch pump is replaced by a 
6X8 inch, the number of strokes will be 
reduced from 300 to less than 75, for the 
same amount of water, or if it makes 75 
strokes it will deliver more water than 
the other did at 300. 



184 



STEAM ENGINEERING 



STEAM AND WATER CYLINDERS 

Attempts hav.e been made to use 
pumps for boiler feeding that were not 
properly proportioned for this service, 
hence all efforts to make them work 
failed because the available force acting 
on the steam piston was not sufficient to 
overcome the resistance opposing the 
water piston. As this was due to igno- 
rance on the subject, or carelessness in 
failing to ascertain the relative diam- 
eters of the steam and water pistons, 
attention is called to it here in order that 
readers may not make the same mistake. 

Fig. 24 is a pimip piston rod with a 
steam piston 10 inches, and a water 



piston 6 inches in diameter on it. As- 
suming that a steam pressure of 100 
pounds to the square inch is realized, 
the total pressure on the steam piston 
is 7,854 pounds, but with the same pres- 
sure per square inch opposing the water 
piston, the total is 2,827 pounds, or a 
difference of 5,027 pounds. 



BOILER FEEDERS 185 

This is graphically illustrated by Fig. 
25 which shows the two pistons together. 




If both of these pistons were the same 
diameter the pump would not work 
because the total resistance would be 
greater than the total force acting on 
the steam piston, for if there is 100 
poimds pressure on a boiler it will usu- 
lally be necessary to pump against at 
least 110 pounds. 

However, it is neither advisable nor 
practicable to use full boiler pressure in 
the steam cylinder of a pump, hence the 
piston must be larger accordingly. 
With 50 pounds pressure on the steam 
piston, the total is 3,927 pounds, and if 
the water piston works against 110 
pounds the total is 3,109 pounds, leav- 
ing a difference of 818 pounds, which is 
enough to cause the pump to run as de- 
sired. This comparison demonstrates 



186 STEAM ENGINEERING 

that much less than boiler pressure will 
drive such a pump, and it also shows 
that the steam piston must always be 
larger than the water piston. 

The next table gives the height to 
which pumps with given steam and 
water pistons or cylinders will elevate 
water. As it is impossible to determine 
the friction that will result from forcing 
water through a system of pipes until 
all conditions are known, or a trial made 
in practice, no attempt is made to in- 
clude it in this table, therefore, the fig- 
ures where the horizontal and the ver- 
tical lines intersect represent the height 
in feet of a colufnn of water acting on the 
water piston that will balance a pressure 
of 50 pounds to the square inch on the 
steam piston. This height may be re- 
duced to pounds by multiplying it by 
.43. 

For illustration, take a steam piston 
12 inches and a water piston 7 inches in 
diameter. By following the horizontal 
line on a level with 12 imtil it inter- 
sects the vertical line under 7, we find 
that the column of water must be 220 
feet high, or 220 X. 43 =94.6 pounds will 
balance 50 pounds on the steam piston. 
It will be noted that wherever the diam- 
eter of the steam piston is twice the 
diameter of the water piston, the col- 
umn of water is 300 feet high. 



BOILER FEEDERS 



187 



2 

u 

"o 

1 

Q 


(M 


CO (N O lOC^ ^ (N 00 

co'*'or^O'*coo 


O 


i-*oo.-iiooor^(Ncoo 

?0Tt(CDI>O^Oxt<O 

r-.^rt(NCO 


o> 


C0u:l05l0Tt^C0lMC0OO 

co-*iot>05cooocoor^ 

r-lr-l(MCOCO 


00 


(Nt>.ioiot>o>oooa>oi 

I-H r-l CM OO CO ■<1< 


I> 


00»OiOOO-*COOO(NOO 
CO»Ot^a2CMiOCMOOTC20 


«o 


Tt<C<IiOCM'-l0500000-^OCO 

coior>oxfi<coooto>oo:i 

.-i.-lrHCMC0-*ir3CD00 


>o 


I>00>OOOI>C<1COO(MOOCMC<J 

co'^t-OTfici-^ocooocDr^ 

T-lrH,-lCMC0^lCI>OJ 


•* 


T-lrHC>^COCO^OCJ 


CO 


iOTt<co'-<o-^<r>CM^ 

1>COIOC^OTJ<C5^00 
rH .-1 C<> CO CO '^ to 00 


CO 


CM '^ 05 O 00 CO >0 CO 

ocoooocot^co 

.-H r-1 CM CO '^ "O CO 00 




g§8§||g 


C^ 


CM CO-* CO C5 


.2 c 
O 


rt o 
w a 


CO-^iOtOt^OOCTJOCMTjHcOOOO 



188 STEAM ENGINEERING 

ORDERING STEAM PUMPS 

Having decided ojn the proper size of 
pump for the maximum quantity of 
water wanted, the engineer should take 
into consideration the following condi- 
tions, and inform the parties of whom 
the pump is to be purchased, concerning 
them, that there may be no misunder- 
standing. 

Whether the water to be pumped is 
hot or cold, or each kind alternately. 
If only cold water is to be used, how 
high must it be lifted by suction, and 
how long is the suction pipe? Is the 
water pure or does it contain impurities 
that must be prevented from passing 
the pump? 

The available steam pressure must be 
considered, also the pressure against 
which water is to be forced. This in- 
cludes consideration of the length of the 
discharge pipe, also the number of short 
turns and valves in it. 

It should be known whether the pump 
is to exhaust into a condenser, the atmos- 
phere, or a heating system of any kind, 
as these conditions effect the back pres- 
sure on the steam piston, or in other 
words they make a difference in the total 
load on the pump. 

When ordering parts for repairs give 
the shop number of the pump, also the 



BOILER FEEDERS 189 

size and serial number if there is one 
provided. The parts wanted are usu- 
ally illustrated in the manufacturer's 
catalogue, therefore one of them should 
be kept on hand for use in an emergency. 
Reading and studying these catalogues 
enables an engineer to thoroughly under- 
stand the machines that he uses. 

DIRECTIONS FOR SETTING AND OPERATING 
STEAM PUMPS 

Having selected a place for the pump 
to be set, have a brick foimdation built 
for it in order to hold it firmly in posi- 
tion. Do not bolt it to a light wooden 
floor that will not be stiff enough to pre- 
vent vibration when new, and that will 
soon rot and become useless. 

Where a pump is to take water under 
pressure from a street main, the suction 
pipe may be made smaller than the suc- 
tion opening calls for, but where water is 
to be lifted from a well, pond or brook, 
it should never be reduced even for a 
low lift, and for a high lift or a long suc- 
tion pipe, it should be made one size 
larger. 

Always make the suction as short and 
straight as conditions will admit, as 
short ells and globe valves add much to 
the necessary friction of water flowing 
through pipes. The highest part of a 



190 STEAM ENGINEERING 

long suction pipe should always be at 
the pump, and the grade should be con- 
tinually downward to the water. If it 
is raised and lowered to suit the ground 
on which it is located, air will collect in 
the high places or air pockets, and cause 
trouble. 

Where it is to be laid under ground, 
cast iron flanged pipe is recommended 
as durable and satisfactory. Special 
care should be taken to keep sand, 
stones, and other foreign matter out of 
the pipe while it is being laid, as but a 
small quantity will cause much trouble 
and annoyance. 

Where it is necessary to locate a valve 
in the straight part of a suction pipe, it 
should be a gate valve in order to create 
as little friction as possible. If there is 
a convenient turn in the pipe, an angle 
valve may be 'used instead of an ell. If 
the suction pipe is long, a foot valve 
should be located on the end of it, then 
by providing a small priming pipe, water 
can be admitted to the suction pipe, 
from an overhead tank, or a street main, 
thus filling the suction pipe and insuring 
prompt operation of the pump when 
steam is turned on. If this suction pipe 
is exposed, provision must be made for 
draining the water out of it when the 
pump is stopped in cold weather. 
Frost plugs in the water cylinder should 



BOILER FEEDERS 



191 



be removed to allow all water to drain 
out. 

A large air chamber should be located 
on the suction pipe in addition to one 
provided on the pump in some cases as 
the latter provides a cushion for the dis- 
charge pipe only. Fig. 26 illustrates 




Fig. ^6 



^ 



192 STEAM ENGINEERING 

the proper location of the air chamber, 
for either one or two suction pipes. If 
two suction pipes are provided; the 
water coming in through them will move 
slowly and quietly. The discharge pipe 
should be of ample size to allow water 
to move at a comparatively slow speed 
when the pump is operated at its full 
capacity. 

When locating the steam pipe, make 
due allowance for expansion when it is 
heated by steam. Put a globe valve 
near the pump and a good double con- 
nection sight feed lubricator above it. 
Before admitting steam to the pump, 
blow out the pipes thoroughly under 
full pressure in order to remove red lead, 
iron chips, and dirt, thus preventing 
such foreign matter from injuring the 
steam cylinder. 

Keep the stuffing boxes well filled 
with first-class packing and do not let it 
remain in use long enough to become 
hard, as it will score the rods, making it 
impossible to keep them from leaking 
steam and water. 

Use as good oil on all pump bearings 
as you would on the best kind of an en- 
gine, and only first-class cylinder oil in 
the steam cylinder. If the pump is to 
remain idle for several days, use an 
extra quantity of cylinder oil during the 
last five minutes that it is run, in order 



BOILER FEEDERS 193 

to keep it from rusting when standing 
still. 

The advantage gained by the use of 
pipes of ample size for both the suction 
and discharge of pumps will be made 
plain by studying the next table, which 
treats pipes from 1 to 6 inches in diam- 
eter. 

G = Gallons discharged per minute. 
V = Velocity of water in feet per 
. second. 

F = Friction of water in pounds per 
square inch for each 100 feet 
of clean iron pipe, or the pres- 
sure required to overcome 
friction. 

It will be noted that as the velocity is 
increased, the friction increases very 
fast, therefore, a constant loss results 
from the use of pipes that are too small 
for the required service, hence while the 
first cost of them is less, they are ulti- 
mately very expensive. 

For illustration, suppose that 20 gal- 
lons of water per minute are forced 
through a 13^ in. pipe, causing a friction 
loss of 1.66 pounds. If the amount of 
water is doubled the loss is increased to 
6.52 pounds. If 35 gallons are forced 
through 1 inch pipe the loss is 37 pounds, 
but 13^ inch pipe will convey the same 
amount with 5.05 pounds loss. 



194 



STEAM ENGINEERING 



FRICTION IN POUNDS PRESSURE PER 

SQUARE INCH 

For each 100 feet of clean iron pipe. G. A. 



.a 

CO 

c 

.c 
c 

.c 

c 

:^ 

c 

S3 

i 
1 


b 


Ellis. C. E. 

«0 O »HC0t^O»fl^{0| 
O ^ t^(NCO»OCO00O> 

6 6 dddddodi 


> 


1.13 

1.70 

2.27 
2.84 
3^40 
3.97 
4.54 
5.11 
5.67 


'fu 


0.09 

0.33 

0.69 

1.22 
1.89 
2.66 
3.65 
4.73 
6.01 
7.43 


> 


1.28 

2.55 

3.83 

5.11 
6.39 
7.66 
8.94 
10.20 
11.50 
12.80 


fo 


0.10 

0.35 
0.74 
1.31 
1.99 
2.85 
3.85 
5.02 
7.76 
11.20 
15.20 
19.50 
25.00 
30.80 


> 


1.13 

2.27 
3.40 
4.54 
5.67 
6.81 
7.94 
9.08 
11.30 
13.60 
15.90 
18.20 
20.40 
22.70 


fu 


d d'-<co'*i>o>(NoJod 


> 


1.63 

3.26 
4.90 
6.53 
8.16 
9.80 
11.40 
13.10 
16.30 
19.60 


fo 


0.12 

0.42 

0.91 

1.60 

2.44 
5.32 
9.46 
14.90 
21.20 
28.10 
37.50 


> 


1.02 

2.04 

3.06 

4.09 

5.11 
7.66 
10.20 
12.80 
15.30 
17.10 
20.40 


(i< 


<Mt>-t^CO(NiO»0<MiOOQO 
^•^CT>COcOt^OiO.-iO-*0 

ddd-H (Nco^ocdoodNoi 

1-HWCO 


> 


0.81 
1.82 
2.73 
3.63 
4.54 
5.45 
6.36 
7.26 
8.17 
9.08 
13.60 
18.20 


iu 


0.31 
1.05 
2.38 
4.07 
6.40 
9.15 
12.04 
16.10 
20.20 
24.90 
56.10 


> 


1.31 
2.61 
3.92 
5.22 
6.53 
7.84 
9.14 
10.40 
11.70 
13.10 
19.60 


> 


dcocDddi^i^'oo 


2.04 
4.08 
6.13 
8.17 
10.20 
12.30 
14.30 
16.30 




O 


lO O "5 O >0 O "O O U5 O IC O lO o o o o o o o o o 
.-<i-Hoaeococo'*><j<ior^o(Niot^oioo»oou50 

_i._(,-(.-4CgCSC0C0"<l*'*»O 



BOILER FEEDERS 195 

RECEIVER PUMPS AS BOILER FEEDERS 

When a receiver is located in the lower 
part of a building that is heated by- 
steam, or in a dry kiln, or in any other 
similar place, all water resulting from 
the condensation of steam flows into 
this receiver, from whence it is taken by 
a pump and returned to the boilers. 
This pump is operated automatically 
by a float in the receiver which rises 
when the water line is raised, and opens 
a balanced valve in the steam pipe of the 
pump, thus starting it and taking out 
the water. When the water level falls, 
the float follows and shutting the bal- 
anced valve stops the pump. If water 
flows continuously to the receiver, the 
float will operate the pump at a proper 
speed to remove it as fast as it comes. 

This is not only an excellent way to 
remove the water of condensation from 
heating systems, but it constitutes a 
very good boiler feeder. 

As there is usually more or less water 
lost through drips, etc., a comparatively 
small quantity of fresh water must be 
put into the boilers every day. It is not 
necessary to run an extra pump for 
• this purpose, for if a cold water pipe is 
tapped into the receiver, or into a con- 
venient return pipe from the heating 
System, it will only be necessary to 



7 



196 STEAM ENGINEERING 




BOILER FEEDERS 



197 



admit cold water as needed and the 
float will do the rest. Fig. 27 illustrates 
a unique device for this purpose. In the 
cast iron receiver there is a float which 
operates the lever shown, and this opens 
and closes the balanced valve on the 
receiver, admitting steam to run the 
pump and return the water to the boiler. 
A double connection lubricator is on the 
steam pipe where it feeds oil through 




198 STEAM ENGINEERING 

the balanced valve, but none of it can 
leak into the receiver and be carried 
into the boilers. A sediment trap is 
provided on the main return pipe which 
prevents foreign matter from going into 
the pump. If this is not provided in 
some form, the water cylinder of the 
pump will probably be badly scored by 
red lead, iron scale, mud, etc., coming 
down from the heating system. 

Fig. 28 is an end view of a pump and 
receiver, showing the steam end of it. 
This has done good service in the 
author's plant for 15 years, and is still in 
good order. A double connection lubri- 
cator is used as shown, except that it is 
not so near the pump, but is located in 
the boiler room, where it can be seen by 
the assistant engineer. All cylinder oil 
passes through the balanced valve thus 
keeping it in good working order, 
but none of the oil can get into the 
receiver. 

Fig. 29 is another end view of the same 
machine, showing the water end of it. 
The outlet is located in the head, thus 
allowing all sediment to stay in the 
receiver where it can do no harm. 

No vent pipe is provided for this re- 
ceiver as it forms part of a closed system 



BOILER FEEDERS 



199 




in which the hot water is not released 
from pressure, but is returned without 
loss of heat. A safety valve is pro- 
vided to protect the receiver from over- 
pressure. 

The following table contains the size 
and capacity of these pumps and receiv- 



^ 



200 



STEAM ENGINEERING 



ers. The third column indicates the 
number of square feet of radiating sur- ; 
face that the pump will drain. 

SIZE AND CAPACITY OF PUMPS AND 
RECEIVERS 



Size of 


Gallons 
mmate 


Square 


Space occupied 


pump 


Length 


Width 


Height 


3x2x3 


12 


5,000 


34 


25 


31 


4^x2^x4 


20 


10,000 


37 


30 


35 


5Mx3J^x5 


35 


20.000 


48 


37 


44 


6x4x6 


60 


40,000 


54 


37 


45 


7Hx5x6 


90 


60,000 


54 


39 


53 


7Hx4HxlOI 90 


60.000 


65 


42 


53 



PUMPS DRIVEN BY ELECTRIC TRANS- 
MISSION OF POWER 

' Fig. 30 illustrates a horizontal power 
pump driven by an electricmotor locat- 
ed at the center of the frame. As a 
high rate of speed at the motor is desira- 
ble while the pump operates at a very 
moderate speed, the reduction is made 
by gears of suitable size. 

This makes a very desirable form of 
boiler feeder wherever electricity is 
available, as the speed can be adjusted 
to meet the requirements of the boilers, 
and the machine can be located in any 
convenient place without regard to con- 



BOILER FEEDERS 



201 



venience in transmitting power as it 
must be considered in other cases, be- 
cause a pair of wires can be carried any- 




where, and a reasonable length of pipe 
does not add much to the cost of opera- 
tion, provided it is of ample size. 



202 



STEAM ENGINEERING 



The following table gives size and ca- 
pacity of these pumps, although larger 
ones are built to order. 



SIZE AND CAPACITY OF ELECTRIC 
PUMPS 



Size in 


Speed 


Gallons 


Gallons 


inches 


per 
minute 


per 
revolution 


per 
minute 


3x6 


60 


.366 


21.96 


3Hx 6 


60 


.498 


29.88 


4x8 


50 


.880 


44.00 


4Hx 8 


50 


1.100 


55.00 


5 xlO 


40 


1.700 


68.00 


6 xl2 


30 


2.920 


87.60 


7 xl2 


30 


4.000 


120.00 


8 xl6 


25 


6.960 


174.80 


9 xl6 


25 


8.800 


220.00 


10 xl6 


25 


10.880 


272.00 


12 xl6 


25 


15.660 


391.50 


14 xl6 


25 


21.320 


533.00 



Fig. 31 illustrates a single acting ver- 
tical triplex pump driven by an electric 
motor geared 5 to 1. The design and 
construction of electric motors have 
been brought so near perfection that any 
kind of current can be used to drive 
them, but when laying out a machine of 
this kind, it is a good idea to take the 
kind available into consideration, espe- 
cially when corresponding with parties 
who make motors, as the proper kind 
must be selected to use the available 
current. 



BOILER FEEDERS 



203 



The next table gives the size and 
capacity of these pumps as ordinarily 
used. It is not good policy to select a 
small electrically driven pump, and 




(r3 



operate it at a higher speed that the list 
calls for, either for the horizontal or 
the vertical types, as vibration and 
wear will be greatly increased by such 
action. 



204 



STEAM ENGINEERING 



SIZE AND CAPACITY OF DIRECT CON- 
NECTED ELECTRIC PUMPS 



Size in 


'Speed 


Gallons 


C^alloos 


inches 


per 

minute 


per 
revolution 


per 
mmute 


2 x2 


70 


.081 


5.67 


21^x2 


70 


.127 


8.89 


2Hx3 


60 


.19 


11.40 


3 x3 


60 


.27 


16.20 


3Hx3 


60 


.37 


22 


3Hx4 


60 


.50 


30 


4 x4 


60 


.65 


39 


4 x6 


60 


.98 


59 


5 x6 


60 


1.53 


91 


5Hx8 


60 


2.46 


147 


6 x8 


55 


2.94 


161 


7 x8 


55. 


4.00 


220 


8 x8 


55 


5.22 


287 


8^x8 


55 


5.90 


324 


9 xlO 


50 


8.26 


413 


10 xlO 


45 


10.20 


459 


11 xl2 


42 


14.81 


622 


12 x|12 


42 


17.62 


740 



INJECTORS 

The fact that a jet of steam when issu- 
ing from a properly proportioned and 
correctly located nozzle, can be used to 
move a body of water from one place to 
another, was known more than 100 
years ago. It was utilized to the extent 
of raising water and allowing it to flow 
into a tank, or other receptacle below 
the apparatus, and later it was found 
practicable to raise water with it into a 
tank located above the apparatus, thus 
forming what we now call an ejector. 



BOILER FEEDERS 205 

In 1858 the eminent French engineer, 
Giffard, discovered that if a jet of steam 
was used to raise water and it was first 
allowed to flow into the air on about a 
level with the apparatus, after working 
in this way until a rapid circulation was 
established, the overflow could be slow- 
ly shut off, when the moving jet of water 
and steam would force itself into the 
boiler against the pressure that was used 
to operate the apparatus, thus giving 
us the injector as we have it in use today. 
To Giffard, then, belongs the honor, not 
of actually inventing the injector, but 
of perfecting it and making a practical 
application of the principles involved 
as we understand them now. 

THEORY OF THE INJECTOR 

Many so-called theories which claim 
to explain the operation of this machine 
have been advanced, some of them by 
men who simply guess at it without 
having knowledge of the natiure of fluids 
and gases, but the following is both 
reasonable and logical. 

The jet of steam used in this machine 
puts into motion a body of water and as 
the steam strikes the water and mingles 
with it, a portion of the great velocity 
of steam flowing into the atmosphere is 
imparted to the mixture which is thus 



206 STEAM ENGINEERING 

thrown forward. If the velocity' of this 
mass is greater than the velocity of 
water flowing out of the boiler under 
given pressure, then the injector will 
feed the boiler, but otherwise it will not. 
This statement applies to ordinary 
injectors designed for boiler feeding. 

In order to intelligently consider this 
matter it is necessary to take into ac- 
count what may properly be called the 
three active parts of an injector. 

First — A nozzle of peculiar shape and 
correct proportions through which steam 
is admitted from the boiler to the 
injector. 

Second — A combining tube or nozzle 
in which steam from the nozzle above 
mentioned and water from the suction 
pipe meet, condensing the steam and 
imparting some of its velocity and the 
resulting force to the Water. 

Third — A delivery nozzle designed 
and proportioned to give the greatest 
possible velocity to the mixture. 

THE VELOCITY OF STEAM 

When considering the first part it be- 
comes necessary to compute the velocity 
of steam when it is allowed to escape 
into the atmosphere, assuming that it 
has the same weight per cubic foot as 
when it was confined in the boiler and 



BOILER FEEDERS 207 

not taking into account the shape of the 
nozzle. The next formula can be used 
for this purpose. 

r 



V 



BX144 
64.32 X = 



W 

B = Boiler pressure as indicated by 
the steam gauge. As this is 
on the square inch it is mul- 
tiplied by 144 to give the 
total pressure per square 
foot. 

W= Weight of one cubic foot of 
steam taken at absolute 
pressure or B + 14.7 pounds. 
V= Velocity in feet per second. 

For illustration take steam at 100 
pounds gauge pressure, the weight of 
which is .2296 pound per cubic foot. 

Substituting these values for the let- 
ters, and making the necessary calcula- 
tion gives the following result: 

) 100 X 144 

/ 64.32 X =2,008.5 

.2296 
feet per second. 



\/e 



VELOCITY OF THE MIXTURE 

When solving this problem the first 
value to determine is the weight of water 
taken up by the injector compared with 



208 STEAM ENGINEERING 

the weight of steam used. It may be 
calculated by the next formula. 

T-H 



H-W 



= P. 



T = Total heat of steam at 'given 

pressure. 
H = Heat units in the water at given 

temperature as it leaves the 

injector. 
W = Heat units in the water at given 

temperature as it enters the 

injector. 
P = Pounds of water taken up by 

the injector for each pound 

of steam used. 

For illustration take steam at 100 
pounds gauge pressure as above men- 
tioned, the total heat of which is 1,178 
above 32 degrees. If water enters the 
injector at 65 degrees each pound of it 
contains 33.01 heat units and if it is 
discharged from the injector at 155 de- 
grees it contains 123.34 heat units. 
Application of the formula gives the 
following result: 

1,178-123.34 _ ^„ , , 

= 11.67 pounds of water 

123.34-33.01 

taken by the injector for each pound of 
steam used, therefore the ratio is 11.67 
to 1. 



BOILER FEEDERS 209 

The velocity of the mixture can now 
be determined by the next formula. 

I = Quantity of steam used which is 
taken at unity or 1. 

P = Poimds of water taken by the 
injector for each pound of 
steam used. 

V= Velocity of steam at given pres- 
sure when flowing into the 
atmosphere. 

M= Velocity of the mixture when 
flowing into the atmosphere. 

This is not the actual velocity of the 
mixture of steam and water entering the 
boiler, because it must overcome boiler 
pressure, which reduces the speed that 
would be attained if there was no re- 
sistance except what is due to atmos- 
pheric pressure. 

Application of this formula to the ex- 
ample stated gives the following result: 

^:^X2,008.5 = 158.67 

feet per second. 

This mixture of steam and water will 
flow into the boiler providing the veloc- 
ity of it as above calculated is greater 



210 STEAM ENGINEERING 

than the velocity of a jet of water would 
be if it was allowed to flow freely from 
the boiler at stated pressure. This 
velocity may be determined by the 
following formula: 



V64.32XH=V. 

, H = Height of a column of water 

whose weight or pressure 
per square inch at the base 
is equal to the boiler pres- 
sure. 

V = Velocity in feet per second. 

The height of this column is found by 
dividing the number of cubic inches in a 
cubic foot by 12 and dividing the quo- 
tient by the weight of water per cubic 
foot at stated temperature. This quo- 
tient is the height of water in feet re- 
quired for one pound pressure per square 
inch, and when this is multiplied by the 
pressure, the product is the total height 
of a column of water the pressure of 
which is equal to the boiler pressure. 

In this case it is 100 pounds by the 
gauge, the temperature of which is 337.8 
degrees. Water at this temperature 
weighs 56.62 pounds per cubic foot. 
Stated as a formula it appears as follows: 

1,728 

- — ^WXP = H. 



I 



BOILER FEEDERS 211 

W= Weight of a cubic foot of water 
at given pressure and tem- 
perature. 

P = Boiler pressure by the gauge. 

H = Height in feet. 

Then, 

1,728 

-7- 56.62 X 100 = 254 feet. 

12 

The velocity with which water would 
flow from this boiler under 100 pounds 
pressure is 

V64.32X254 = 127.82 feet per second. 

The difference between the velocity 
of the mixture coming from the delivery 
tube, and water issuing from the boiler 
is 158.67-127.82 = 30.85 feet per sec- 
ond, which is sufficient to account for 
the action of the injector. 

Gauge pressure is used in these calcu- 
lations because both steam and water 
are supposed to be flowing into the 
atmosphere. If they were flowing into 
a vacuum it would be necessary to use 
absolute pressure. 

When calculating these results it is 
necessary to know the weight of water 
per cubic foot under various pressures 
with corresponding temperatures. This 
will be foimd in the next table, from zero 
by the gauge to 135 pounds pressure. 



212 



STEAM ENGINEERING 



WEIGHT OF WATER PER CUBIC FOOT 
ABOVE 2120 FAH. 



Pres- Temper- 
sure ature 




0.3125 
1 

3 
4 
5 



212 

213.04 

216.33 

219.45 

222.40 

225.25 

227.95 

230.60 

233.10 

235.49 

237.81 

240.07 

242.24 

244.32 

246.35 

248.33 

250.26 

252.13 

253.98 

255.77 

257.52 

259.22 

260.88 

262.50 

264.09 

265.65 

267.17 

268.66 

270.12 

271.55 

272.96 

274.33 

275.68 

277.01 

278.32 

279.62 

280.89 

282.14 

283.39 

284.58 

285.76 

286.96 

288.06 

289.24 



Weight 


^ Pres- 
sure 


Temper- 
ature 


59.83 


43 


290.37 


59.81 


44 


291.48 


59.73 


45 


292.58 


59.65 


46 


293.66 


59.59 


47 


294.73 


59.52 


48 


295.78 


59.45 


'49 


296.82 


59.38 


,50 


297.84 


59.32 


51 


298.85 


59.27 


52 


299.85 


59.21 


53 


300.84 


59.15 


64 


301.81 


59.09 


55 


302.77 


59.05 


56 


303.72 


59.00 


57 


304.69 


58.95 


58 


305.60 


58.90 


59 


306.52 


58.85 


60 


307.42 


58.80 


61 


308.38 


58.75 


62 


309.22 


58.71 


63 


310.11 


58.67 


64 


310.99 


58.63 


65 


311.86 


58.59 


66 


312.72 


58.56 


67 


313.57 


58.51 


68 


314.42 


58.48 


69 


315.25 


58.43 


70 


316.08 


58.40 


71 


316.90 


58.36 


72 


317.71 


58.33 


73 


318.51 


58.29 


74 


319.31 


58.26 


75 


320.10 


58.23 


76 


320.88 


58.19 


77 


321.66 


58.16 


78 


322.42 


58.13 


79 


323.18 


58.09 


80 


323.94 


58.06 


81 


324.67 


58.03 


82 


325.43 


58.00 


83 


326.17 


57.97 


84 


326.90 


57.94 


85 


327.63 


57.91 


86 


328.35 



■ Weight 



BOILER FEEDERS 



213 



Weight of Water per Cubic Foot Above 212* 
Fah. 



Pres- 
sure 


Temper- 
ature 


Weight 


Pres- 
sure 


Temper- 
ature 


Weight 


87 


329.07 


56.86 


98 


336.58 


56.66 


88 


329.78 


56.84 


99 


337.23 


56.64 


89 


330.48 


56.82 


100 


337.89 


56.62 


90 


331.18 


56.80 


105 


341.00 


56.54 


91 


331.87 


56.79 


110 


344.1 


56.46 


92 


332.56 


56.72 


115 


347.1 


56.38 


93 


333.24 


56.75 


120 


350.0 


56.30 


94 


333.92 


56.73 


125 


352.8 


56.22^ 


95 


334.59 


56.71 


130 


355.6 


56.14 


96 


335.26 


56.69 


135 


358.4 


56.07 



EFFICIENCY OF THE INJECTOR 

Engineers who take pride in clainiing 
to be practical men only, seem to be 
still further satisfied with their position 
when told that theoretically an injector 
is the most efficient boiler feeder known 
to modem science. This conclusion 
on their part is due to the fact that in 
practice more fuel is required to main- 
tain a given steam pressure when feed- 
ing with one of these. very simple ma- 
chines, than when using a direct* acting 
steam pump, or one of the many forms 
of power pumps, hence it does not seem 
to be ah efficient machine. 

While this appears to be a conclusive 
argument, it is not really so. Experi- 
ments made by Prof. Carpenter show 
that when considered only as a pump. 



214 STEAM ENGINEERING 

an injector has a very low efficiency 
when compared with other kinds, as it 
developed from 161,000 to 2,752,000 
foot pounds of work for each 1,000 
pounds of steam used, or its equivalent 
in burning 100 pounds of coal. 

As small direct acting steam pumps, 
which are known to be very wasteful, 
develop from 4,000,000 to 8,000,000 foot 
pounds on the same basis, the difference 
is apparent, and when compared with 
high grade pumping engines which de- 
velop from 100,000,000 to 140,000,000 
foot pounds, ■ the theoretical rating of 
injectors when considered as pumps 
only, agrees very closely with the prac- 
tical engineer's idea of their performance 
in every day use. 

The fact remains, however, that the 
injector is a very efficient boiler feeder, 
because all heat in the steam used 
(except a small part lost by radiation) 
is returned to the boiler in the feed 
water. Its seeming inefficiency is due 
to the fact that live steam is used for the 
entire process, while with pumps the 
heating is usually done with exhaust 
steam that is a waste product. In 
other words an injector alone is com- 
pared with a pump and a heater oper- 
ating in combination, which is evidently 
unfair. 



BOILER FEEDERS 215 

CAPACITY OF INJECTORS 

The maximum capacity of injectors 
is given in the manufacturers' catalogue, 
but this does not necessarily mean that 
they will deliver the given amount of 
water under all conditions, therefore, 
when selecting an injector to feed one or 
more boilers, it is a good idea to decide 
on the kind wanted, then select one that 
will deliver fully as much as can be used 
under stated conditions. 

Put a globe valve in the suction pipe 
and regulate the supply by it to meet 
the varying demands. 

If a valve in the suction pipe can be 
used to regulate the supply, it is plain 
that if the pipe is lengthened it will 
affect the amount of water brought to 
the injector, hence this must be taken 
into consideration. A careful test was 
made on a certain injector that deliv- 
ered 108 cubic feet of water per hour 
when lifting it one foot. The lift was 
increased one foot for each trial until it 
measured 8 feet, then 2 feet for each 
until it was 20 feet, but the steam pres- 
sure was carefully kept uniform during 
the whole test. 

The increased lift reduced the capac- 
ity to about 52 cubic feet per hour. As 
these trials were made in rapid succes- 
sion and the capacity was less every 



216 STEAM ENGINEERING 

time that the lift was increased, the 
reduction could not be charged to any 
other change. As this shows a reduc- 
tion of more than 50 per cent, it may 
not be practicable to depend on so great 
a change in every day practice, but a 
reduction of 40 per cent, can be made 
without affecting the reliability of the 
machine as a boiler feeder, and that 
should be considered sufficient. 

It may not be possible to do this with 
every injector in the market, but where 
it is impossible to thus reduce the capac- 
ity, provided the water is not warmer 
than 70 degrees Fah., it is evidence of an 
inferior machine. If a greater varia- 
tion than this is required in any special 
case, two injectors should be installed, 
thus giving a range of from 60 per cent, 
of the capacity of one to the full capac- 
ity of both, or in other words this plan 
allows a reduction of 70 per cent, from 
the full capacity. 

When the manufacturer of an injector 
publishes a list of sizes and the capacity 
of each size, the given quantities are 
usually based on the use of a stated 
steam pressure. If a higher pressure is 
carried in service, the capacity will be 
increased, and where it is lower the 
amount delivered will be less accord- 
ingly. 



BOILER FEEDERS 



217 



TYPES OF INJECTORS 

Fig. 32 represents a type of injector 
[that was considered very good until 




better ones were invented. It would 
lift water and deliver it to a boiler, but 



218 



STEAM ENGINEERING 



it required adjusting for every change 
in steam pressure, which is not consid- 
ered good practice now. If too much 
steam was admitted some of it would 
escape at the overflow, and if the supply 
of water was too great, a portion of it 
came out of the same place, thus show- 
ing the need of attention in either case. 
Fig. 33 is an up-to-date automatic 
injector. It requires no adjustment 




BOILER FEEDERS 



219 



I for varying steam pressure, and when 
lifting water it is not necessary to give 



0=0 




Fi,c.34- 



220 



STEAM ENGINEERING 



this part any attention unless too much 
is taken up, when the supply may be 
reduced. 

Fig. 34 represents one kind of a Han- 
cock inspirator, which is another name 
for a double tube injector. One set of 
tubes lifts the water and delivers it to 
the other which forces it into the boiler. 
Although it has three valves to operate 
it is not a complicated machine, and as 
each valve must be handled separately, 
it is not difficult to locate a defect which 




BOILER FEEDERS 221 

prevents smooth operation on account 
of this feature. 

Fig. 35 represents a double tube in- 
jector in which all valves are connected 
and are operated by one lever, therefore 
the starting and stopping process con- 
sists of one operation only, so far as the 
injector is concerned. 

Fig. 36 is an exhaust steam injector 
that will feed a boiler against 75 pounds 
pressure, without using live steam. If 
more than 75 pounds are carried, a 
small quantity of live steam must be 
used. In the former case there can be 
no doubt of the economy of its perform- 
ance, as only a waste product is used to 
heat the feed water and force it into the 
boiler. This statement applies to cases 
where exhaust steam is not used for 
heating or other useful purpose, and 
even then it is a superior machine, 
because it uses only what is actually 
necessary, allowing the remainder to go 
where it is wanted. It is not necessary 
to create back pressure in the exhaust 
pipe, to operate this injector, but the 
piping must be arranged so as to afiEord 
an ample supply directly to the injector. 
Do not arrange the piping so that ex- 
haust steam will rush through a straight 
pipe in which there is a tee with a side 
outlet through which to supply the in- 
jector, as it will not always work well 



222 STEAM ENGINEERING 




FIS.S(^ 



BOILER FEEDERS 223 

unless there is more or less back pres- 
sure. Let the side outlet take away 
what steam the injector does not use but 
the direct supply should come through 
the straight line. 

PIPING INJECTORS 

Fig. 37 illustrates a good plan for 
piping an injector, where water is taken 
from a street main or an elevated tank, 
especially if the same supply pipe is used 
for other purposes. If such a pipe de- 
livers water directly to an injector under 
pressure it will work well if nothing is 
done to change the pressure, but if water 
is drawn off at other places along the line 
sudden opening and closing of the cocks 
of valves will quickly change the pres- 
sure, and as a general rule will cause the 
injector to "kick" or stop working. 

If there is air confined in such a pipe 
it will cause trouble when it comes to 
the injector, because water cannot con- 
dense it and thus reduce its volume, 
consequently the injector stops, and if 
the engineer or fireman does not discov- 
er the failure until the water line in the 
boiler falls too low, the trouble may be 
serious. Steam will be blown into the 
water pipes inider such conditions, and 
carry heat where it is not wanted. 



224 



STEAM ENGINEERING 



The illustration shows a water main 

2 located under the floor with an outlet 

3 that delivers water through an angle 
valve 4 and a float valve' 5 to a tank. 
This arrangement of piping and valves 
allows water to flow freely at the proper 
rate to secure an even supply. Where 
attempts are made to regulate the sup- 
ply by hand, using a valve like 4 or any 
similar arrangement without the float 
or tank valve, the water level is sure to 




be too high or too low sooner or later, 
and cause trouble. 

When water is no longer wanted it 
should be shut off at 4, as the automatic 



BOILER FEEDERS 



225 



valve 5 is intended to regulate the sup- 
ply while the injector is in use. If 




water leaks into this tank when not in 
use, it passes out through the overflow- 
pipe 6. 

A Hancock Inspirator is shown in this 



226 STEAM ENGINEERING 

particular case, but any other kind of an 
injector should be piped in the same 
way. If for any reason the injector is 
heated by steam leaking through it 
when not in use, cold water from the 
main 2 can be admitted through 7 and 
cool it, also the suction pipe without 
spilling water on the floor, as it passes 
through 8 into the tank. This may also 
be used to supply the injector when the 
steam pressure is too low to take water 
from the tank. 

Under ordinary conditions, to start 
the machine, admit steam at 9 drawing 
water through the suction pipe 10 
which is fitted with an angle check valve. 
(This does not assist the injector, but it 
prevents water from flowing into the 
tank through 10 when admitted at 7.) 

After water has circulated rapidly 
through the pipes for a few minutes, 
and there is no more hot water in the 
steam pipe, close 11, open 12 one-quarter 
of a turn, and slowly close 13 with a steady 
movement, as none of the valves used to 
control an injector of any kind should 
be operated with an erratic motion. 

This will cause hot water to flow 
through 14 to the boilers. If the full 
capacity is not required, the supply may 
be regulated by the angle valve 15. 
When not in use, valves 11 and 13 should 
be left open and 12 closed. 



SECTION 3 
STEAM ENGINES 

DEFINITIONS 

Various terms which are employed in 
the descriptions of steam engines are 
used in such different ways by engineers 
and others, that it is often difficult to 
decide definitely what is meant unless 
many words are added in explanation. 
As this leads to misunderstandings, the 
practice should be reformed and put on 
a uniform basis. 

For illustration of this point it is only 
necessary to call attention to the fact 
that formerly when an engineer men- 
tioned a "slide valve engine" we knew 
that he meant an engine that was fitted 
with a D slide valve which always cut 
off steam at one point in the stroke, and 
the speed was regulated by a throttling 
governor. 

At the present time of writing he may 
mean that same kind of an engine, or he 
may mean one that is fitted with a bal- 
anced valve working between the cylin- 
der and a pressure plate. It cuts off 
steam automatically according to the 
load on the engine, and the speed is reg- 
ulated by a shaft governor. 



228 STEAM ENGINEERING 

The following definitions are suggested 
as being concise and expressive when 
properly used. 

SLIDE VALVE ENGINE 

An engine in which steam is admitted 
to the cylinder, cut-off and released by 
an unbalanced D slide valve, operated 
by direct connection to a fixed eccentric, 
therefore the point of cut- off is constant 
under all conditions. The speed is reg- 
ulated by a throttling governor which 
wiredraws the steam and thus varies 
the mean effective pressure to suit 
changes in the load. 

AUTOMATIC ENGINES 

An engine in which steam is admitted 
to the cylinder at little less than boilei: 
pressure under good conditions, cut-off 
at varying points according to the load 
carried, thus regulating the speed, after 
which it is released and passes to the ^ 
exhaust pipe. Such an engine may have 
four valves, or it may have but one, and 
it may be run at a low or a high speed, 
but neither the number of valves nor 
the speed at which it operates, has any ■ 
effect on the principles underlying its 
operation, hence do not form a basis for 
the name given. 



STEAM ENGINES 229 

LOW SPEED AUTOMATIC FOUR VALVE 
ENGINE 

An engine designed for a low rotative 
speed, but not necessarily a low piston 
speed. Steam is admitted through two 
ports to the cylinder at the proper time, 
by two steam valves which also cut off 
the supply according to the load carried. 
This steam is exhausted through two 
separate ports by two exhaust valves. 
The Corliss is the best known member 
of this class, but it also includes many 
others. The speed is usually regulated 
by a fly ball governor, but there are ex- 
ceptions to this general rule. 

HIGH SPEED AUTOMATIC FOUR VALVE 
ENGINE 

An engine that is operated at a high 
rotative speed which generally insures 
a comparatively high piston speed. 
Steam is admitted to the cylinder by two 
separate valves which also cut off the 
supply when the' proper amount has been 
admitted. Steam is exhausted after its 
work is done, by two other separate 
valves. The speed is controlled by a 
shaft governor. 



230 STEAM ENGINEERING 

HIGH SPEED AUTOMATIC SINGLE VALVE 
ENGINE 

This engine is run at a high speed, 
and the steam is admitted to the cylin- 
der, cut off and released by a single 
valve, which may be either of the flat or 
of the piston type. The speed is regu- 
lated by a shaft governor, 

SINGLE ACTING ENGINE 

In this type steam is admitted to one 
end only of the cylinder, and duly ex- 
hausted from it. It may be horizontal 
or vertical or a combination of the two 
kinds. If there are two cylinders, each 
taking steam directly from the boiler, 
and operating independently, it is a 
simple engine, but if steam is admitted to 
one cylinder directly from the boiler, 
and from thence is exhausted into an- 
other, it is a compound engine. The 
number of cylinders does not affect the 
name, as it is determined by the way in 
which steam is used. 

DOUBLE ACTING ENGINE 

Steam is admitted to both ends of the 
cylinder alternately, and after doing its 
work it is exhausted indefpendently. 
The number of cylinders or of valves 



STEAM ENGINES 231 

used does not determine its name 
neither does the speed at which it is 
operated, nor the means adopted for 
controlling the speed. In the absence 
of a statement to the contrary, every 
engine is assumed to be double acting. 

SINGLE ENGINE 

Any kind of a steam engine that is 
fitted with one cylinder and one crank. 
Neither the way in which steam is con- 
troUedy nor the speed at which it is oper- 
ated are responsible for the name given. 

DOUBLE ENGINE 

Any kind of a steam engine that is 
fitted with two cylinders and two cranks, 
each operating entirely independently 
of the other. 

SIMPLE ENGINE 

An engine in which steam is admitted 
directly from the boiler to one cylinder, 
after which it is exhausted to the atmos- 
phere, or to some form of condenser. 

COMPOUND ENGINE 

An engine in which steam is admitted 
directly from the boiler to one cylinder, 
and after its work here is done it is ex- 



232 STEAM ENGINEERING 

hausted to another much larger in diam- 
eter where it is still further expanded 
and caused to do useful work. 

TRIPLE EXPANSION ENGINE 

In this kind of an engine steam is ad- 
mitted to one cylinder, then exhausted 
into another of greater diameter, after 
which it passes to another still larger, 
where it is further expanded and then 
exhausted to the atmosphere or to some 
form of condenser. In some cases this 
would make the third cylinder and pis- 
ton very large and heavy, hence to over- 
come this objection two cylinders are 
provided, the combined area of which is 
equal to the area of one of suitable size, 
and one-half of the steam that comes 
from the second or intermediate cylin- 
der, passes to each of these, and is duly 
expanded to a low pressure. Although 
such an engine has four cylinders, it is 
still a triple expansion engine. 

NON-CONDENSING ENGINE 

Any kind of an engine, or a combina- 
tion of two or more of the above men- 
tioned kinds, in which steam is exhaust- 
ed into the atmosphere after performing 
the work for which it was generated. It 
is sometimes called a "high pressure 



STEAM ENGINES 233 

engine," but this is not correct because 
the term has no definite meaning at the 
present time. 

CONDENSING ENGINE 

An engine in which steam is used in 
one or more cylinders, then exhausted 
into a partial vacuum, and thence into a 
condenser. It was formerly called a 
"low pressure engine," but the term is 
obsolete and should not be used, because 
many condensing engines are rim under 
very high pressures. 

SELECTING AN ENGINE 

When a steam engine is to be installed 
in a given place, there are many points 
to be taken into consideration before an 
intelligent conclusion can be reached 
as to the kind and size that will give best 
results in that particular place. 

If for any reason skilled labor to oper- 
ate it is not available, either on account of 
unfavorable conditions in regard to loca- 
tion, or because the owners have decided 
in advance that they will not pay an 
engineer what a good one is worth, the 
simplest and cheapest engine should be 
purchased, for an economical machine 
is of necessity more or less complicated, 
or other conditions are its equivalent, 



234 STEAM ENGINEERING 

and the best one that can be made will 
prove unsatisfactory if not taken care of 
by a competent engineer. 

Where fuel is expensive, the cost of it 
becomes an important item in the ex- 
pense account, and failure to properly 
manage this part of a business has caused 
more than one failure where competition 
is sharp and margins are small, and the 
same blunder will cause many more in 
the future. The problem will not be 
solved in any case by purchasing a com- 
plicated engine and then hiring a laborer 
without qualifications for this important 
work to take charge of it, expecting to 
get a competent engineer to examine 
the plant once or twice each year, and to 
correct any mismanagement that may 
be found. These matters need constant 
intelligent care and nothing else will 
answer the purpose. 

For some places the short stroke, high 
speed engine is the most appropriate, 
but for others it is not suitable, and it is 
not possible to mention all points bear- 
ing on the subject in a treatise that can 
only grasp the general outlines of it. 
As the low speed engine of two or three 
decades ago is out of the question, a 
medium between the two extremes will 
be considered, or in other words an en- 
gine in which the diameter of the cylin- 
der is about equal to one-half of the 



STEAM ENGINES 235 

stroke, and the piston speed is approxi- 
mately 600 feet per minute. 

For the purpose of illustrating several 
points it is assumed that 300 indicated 
horse power is wanted, and that a simple 
non-condensing automatic engine will be 
used. The boiler pressure is limited to 
110 pounds by the gauge, with a reduc- 
tion of 5 pounds for the initial pressure 
in the cylinder. Under ordinary condi- 
tions the mean effective pressure of a 
non-condensing engine may be taken* at 
about one-half of the boiler pressure by 
the gauge, or a few pounds less than this 
indicates, for economical results in prac- 
tice. 

From this explanation it will be plain 
that when estimating the size of an en- 
gine that will be required for a given case 
the horse power of it is a fixed, quantity 
the piston speed is approximately deter- 
mined by economical considerations, 
also the mean effective pressure in con- 
nection with the available boiler pres- 
sure. It should be remembered that 
this is only an estimate and not a close 
calculation. This leaves only the diam- 
eter of the cylinder to be determined, 
aad the rule for this purpose may be 
stated as follows: 

Multiply the horse power required by 
the number of foot pounds constituting 
one horse power. Divide the product 



^ 



236 STEAM ENGINEERING 

by the piston speed in feet per minute, 
multiplied by the mean effective pres- 
sure. Divide the quotient by .7854 and 
extract the square root, which is the re- 
quired diameter. It may be stated as a 
formula as follows: 



V— -— -7- .7854 = D. 

SXP 

H-P = Horse power required. 

F-P = Ft. pounds for one horse power. 

S = Piston speed in feet per minute. 

P = mean effective pressure. 

D = Diameter of the cylinder, or 
piston. 

In this case the value of H-P is 300 
while S equals 600 and P is taken at 60 
for reasons previously explained, and 
the value of F-P is always 33,000. Sub- 
stituting these values for the letters 
gives the following result: 

/300X33,000 __ _^ . , 

V '■ ^ .7854 = 20.5 inches, 

600X50 ' 

which in practice would be called 20 in. 
As the stroke is to be about twice the 
diameter of the cylinder, it will be fixed 
at 42 inches, or 7 feet for each revolution 
therefore, in order to secure a piston 
speed of 600 feet the fly wheel must re- 
volve nearly 90 times per minute. 



STEAM ENGINES 237 

Placing these results in concise form, 
the following data appears: 

Boiler pressure, 110 pounds 

Initial pressure, 105 " 

Mean effective pressure, 50 " 

Diameter of cylinder, 20 inches^j 

Length of stroke, 42 " 

Revolutions per minute, 90 

The diameter of this cylinder may^be 
determined by a shorter method which 
does not include the extraction of square 
root. Omitting this part it appears as 
follows: 

300X33,000 



600X50 



-=330 



Taking a table of the areas of circles 
and following down the column of areas 
imtil 330 or the nearest value to it is 
found, the corresponding diameter is 
20^ inches, which is called 20 as this 
is only an estimate of the requirements. 

In order to prove the value of this 
estimate, it becomes necessary to deter- 
mine the power that this engine will de- 
velop imder these conditions, and the 
following rule can be used for this pur- 
pose. 

Multiply the area of piston by the 
mean effective pressure, and by the pis- 



238 STEAM ENGINEERING 

ton speed in feet per minute. Divide 
the product by 33,000. 

As some engineers prefer such a rule 
stated as a formula, it is herewith given: 



33,000 



A = Area of piston in square inches. 

P = Mean effective pressure. 

S = Piston speed in feet per minute, 
or the stroke in feed X2X 
the number of revolutions 
per minute. 

I. H. P. = Indicated horse power. 

Applying this formula to the example 
given to illustrate a rule for estimating 
the^size of an engine required for a 
given case, gives the following result: 



314.16X50X630 ^^^ ^„ 

=299.88 

33,000 



indicated horse power, which shows a 
very close approximation or estimate. 

As a table of the areas of circles is very 
convenient for use in this connection, it 
is given on the following pages for this 
purpose. 



STEAM ENGINES 



239 



CIRCUMFERENCES AND AREAS OF 
CIRCLES 

Advancing by Eighths 



Diana. 


Circum. 


Area 


Diam. 


Circum. 


Area 


1 


3.1416 


.7854 


3H 


11.781 


11.045 


lV4e 


3.3379 


.8866 


sAe 


11.977 


11.416 


IH 


3.5343 


.9940 


ZVi 


12.174 


11.793 


1%6 


3.7306 


1.1075 


31%6 


12.370 


12.177 


IM 


3.9270 


1.2272 


4 


12.566 


12.566 


1%6 


4.1233 


1.3530 


4Vl6 


12.763 


12.962 


IH 


4.3197 


1.4849 


4H 


12.959 


13.364 


l%e 


4.5160 


1 . 6230 


43/16 


13.155 


13.772 


IH 


4.7124 


1.7671 


4M 


13.352 


14.186 


me 


4.9087 


1.9175 


45A6 


13.548 


14.607 


IVa 


5.1051 


2.0739 


4H 


13.744 


15.033 


li^c 


5.3014 


2.2365 


4^/16 


13.941 


15.466 


IH 


5.4978 


2.4053 


4M 


14.137 


15.904 


113/16 


0.6941 


2.5802 


4%6 


14.334 


16.349 


1% 


5.8905 


2.7612 


4H 


14,530 


16.800 


11%6 


6.0868 


2.9483 


4m6 


14.726 


17.257 


2 


6.2832 


3.1416 


4^ 


14.923 


17.721 


2^6 


6.4795 


3.3410 


4^6 


15.119 


18.190 


2H 


6.6759 


3.5466 


4?^ 


15.315 


18.665 




6.8722 


3.7583 


4"/l6 


15.512 


19.147 


2Ji 


7.0686 


3.9761 


5 


15.708 


19.635 
20.129 


26^6 


7.2649 


4.2000 


6Vl6 


15.904 


2^ 


7.4613 


4.4301 


6H 


16.101 


20.629 


2T.i6 


7.6576 


4.6664 


58A6 


16.297 


21.135 


21^ 


7.8540 


4.9087 


bH 


16.493 


21.648 


2%6 


8.0503 


5.1572 


5%6 


16.690 


22.166 


2fl 


8.2467 


5.4119 


bH 


16.886 


22.691 


21^6 


8.4430 


5.6727 


bMa 


17.082 


23.221 


2^ 


8.6394 


5.9396 


5H 


17.279 


23.758 


2i8Ae 


8.8357 


6.2126 


5%6 


17.475 


24.301 


2^^ 


9.0321 


6.4918 


bVz 


17.671 


24.850 


2i'H6 


9.2284 


6.7771 


51^6 


17.868 


25.406 


3 


9.4248 


7.0686 


bH 


18.064 


25.967 


3^6 


9.6211 


7.3662 


51%6 


18.261 


26.535 


SH 


9.8175 


7.6699 


bVi 


18.457 


27.109 


3?i6 


10.014 


7.9798 


5»^l6 


18.653 


27.688 


3K 


10.210 


8.2958 


6 


18.850 


28.274 


3%6 


10.407 


8.6179 


6H 


19.242 


29.465 


3^ 


10.603 


8.9462 


6Ji 


19.635 


30.680 


3%6 


10.799 


9.2806 


6H 


20.028 


31.919 


3^ 


10.996 


9.6211 


6H 


20.420 


33.183 


3%6 


11.192 


9.9678 


6H 


20.813 


34.472 


3H 


11.388 


10.321 


&H 


21.206 


35.785 


31^16 


11.585 


10.680 


6H 


21.598 


37.122 



240 STEAM ENGINEERING 

Circiunferences and Areas of Circles — Continued 



Diam. 


Circum. 


Area 


Diam. 


Circum 


Area 




7 


21.991 


38.485 


12 H 


40.448 


130.19 


; 


7M 


22.384 


39.871 


13 


40.841 


132.73 


; 


7H 


22.776 


41.282 


13H 


41.233 


135.30 


1 


7|/g 


23.169 


42.718 


13M 


41.626 


137.89 




7H 


23.562 


44.179 


13^ 


42.019 


140.50 




7H 


23.955 


45.664 


13H 


42.412 


143.14 




78^ 


24.347 


47.173 


13H 


42.804 


146.80 




7^ 


24.740 


48.707 


13% 


43.197 


148.49 




8 


25.133 
25.525 


50.265 


13 H 


43.590 


151.20 




8J4 


51.849 


14 


43.982 


153.94 




SH 


25.918 


53.456 


14 H 


44.375 


156.70 




SH 


26.311 


55.088 


14^ 


44.768 


159.48 




SVz 


26.704 


56.745 


14 H 


45.160 


162.30 




SH 


27.096 


58.426 


141^ 


45.553 


165.13 




SH 


27.489 


60.132 


14 H 


45.946 


167.99 




8H 


27.882 


61.862 


UH 


46.338 


170.87 ! 




9 


28.274 


63.617 


UPi 


46.731 


173.78 




9H 


28.667 


65.397 


15 


47.124 


176.71 




9H 


29.060 


67.201 


15 M 


47.517 


179.67 




9H 


29.452 


69.029 


15 M 


47.909 


182.65 ! 




9H 


29.845 


70.882 


15 H 


48.302 


185.66 




9H 


30.238 


72.760 


IbVi 


48.695 


188.69 




9H 


30.631 


74.662 


15 H 


49.087 


191.75 




9Vi 


31.023 


76.589 


15% 


49.480 


194.83 




10 


31.416 


"8.540 


15 >^ 


49.873 


197.93 




10 M 


31.809 


80.516 


16 


50.265 


201.06 




10 K 


32.201 


82.516 


16>^ 


50.658 


204.22 i 




10 H 


32.594 


84.541 


IQK 


51.051 


207.39 1 




101^ 


32.987 


86.590 


16H 


51.444 


210.60 , 




10 H 


33.379 


88.664 


163^ 


51.836 


213.82 




10^ 


33.772 


90.763 


16H 


52.229 


217.08 i 




10 Vi 


34.165 


92.886 


16% 


52.622 


220.35 




11 


34.558 


95.033 


16^ 


53.014 


223.65 




ll>i 


34.950 


97.205 


17 


53.407 


226.98 




11^ 


35.343 


99.402 


17 M 


53 ..800 


230.33 




IIH 


35.736 


101.62 


17 M 


54.192 


233.71 




UVi 


36.128 


103.87 


17 H 


54.585 


237.10 




IIH 


36.521 


106.14 


17 M 


54.978 


240.53 




IIM 


36.914 


108.43 


17 H 


55.371 


243.98 




11^ 


37.306 


110.75 


17 H 


55.763 


247.45 




12 


37.699 


113.10 


nvi 


56.156 


250.95 




12^ 


38.092 


115.47 


18 


56.549 


25*. 47 




12 >i 


38.485 


117.86 


18 H 


56.941 


258.02 




12^ 


38.877 


120.28 


18% 


57.334 


261.59 




12 >^ 


39.270 


122.72 


181^ 


57.727 


265.18 




125^ 


39.663 


125.19 


18^ 


58.119 


268.80 




125^ 


40.055 


127.68 


18 H 


58.512 


272.45 





STEAM ENGINES 241 

Xircumferences and Areas of Circles — Continued 



Diaxn. 


Circmn. 


Area 


Diam. 


Circum. 


Area 


18M 


58.905 


276.12 


24 H 


77.362 


476.26 


18?^ 


59.298 


279.81 


24^ 


77.754 


481.11 


19 


59.690 


283.53 


24 4 


78.147 


485.98 


19 >^ 


60.083 


287.27 


25 


78.540 


490.87 


19 H 


60.476 


291.04 


25 H 


78.933 


495.79 


19 H 


60.868 


294.831 


25 H 


79.325 


500.74 


19^ 


61.261 


298.651 


25 H 


79.718 


505.71 


191^ 


61.654 


302.49 


2514 


80.111 


510.71 


19^ 


62.046 


306.35 


25 SA 


80.503 


515.72 


19^ 


62.439 


310.24 


25 M 


80.896 


520.77 


20 


62.832 


314.16 


25 7A 


81.289 


525.84 


20H 


63.225 


318.10 


26 


81.681 


530.93 


20 M 


63.617 


322.06 


26 H 


82.074 


536.05 


20 H 


64.010 


326.05 


26 M 


82.467 


541.19 


201^ 


64.403 


330.06 


26 H 


82.860 


546.35 


20 H 


64.795 


334.10 


26 >i 


83.252 


551.55 


20 M 


65.188 


338.16 


26H 


83.645 


556.76 


20 H 


65.581 342.25 


26^ 


84.038 


562.00 


21 


65.973 


346.36 


262^ 


84.430 


567.27 


21K 


66.366 


350.50 


27 


84.823 


572.56 


21>i 


66.759 


354.66 


27 M 


85.216 


577.87 


21^ 


67.152 


358.84 


21 K 


85.608 


583.21 


21H 


67.544 


363.05 


27 H 


86.001 


588.57 


21H 


67.937 


367.28 


27 J^ 


86.394 


593.96 


21?i 


68.330 


371.54 


27^ 


86.786 


599.37 


21^ 


68.722 


375.83 


27 M 


87.179 


604.81 


22 


69.115 


380.13 


27^ 


87.572 


610.27 


22 H 


69.508 


384.46 


28 


87.965 


615.75 


22 M 


69.900 


388.82 


283^ 


88.357 


621.26 


22 H 


70.293 


393.20 


28 M 


88.750 


626.80 


22 H 


70.686 


397.61 


28 H 


89.143 


632.36 


22^ 


71.079 


402.04 


28 M 


89.535 


637.94 


22 H 


71.471 


406.49 


28 H 


89.928 


643.55 


22^ 


71.864 


410.97 


28 M 


90.321 


649.18 


23 


72.257 


415.48 


28 J^ 


90.713 


654.84 


23 H 


72.649 


420.00 


29 


91.106 


660.52 


iS 


73.042 


424.56 


29 H 


91.499 


666.23 


73.435 


429.13 


29 K 


91.892 


671.96 


231^ 


73.827 


433.74 


29 H 


92.284 


677.71 


23 H 


74.220 


438.36 


29 H 


92.677 


683.49 


23^ 


74.613 


443.01 


29^ 


93.070 


689.30 


& 


75.006 


447.69 


29 5i 


93.462 


695.13 


54 H 


75.398 


452.39 


29 J^ 


93.855 


700.98 


75.791 


457.11 


30 


94.248 


706.86 


24 >i 


76.184 


461.86 


30 M 


94.640 


712.76 


24 H 


76.576 


466.64 


30 Ji 


95.033 


718.69 


24 H 


76.969 


471.44 


30H 1 95.426 


724.64 



242 STEAM ENGINEERING 

Circumferences and Areas of Circles — Continued 



Diam. 


Circum. 


Area 


Diam. 


Circum. 


Area 


30 J^ 


95.819 


730.62 


36H 114.275 1039.2 


30 5^ 


96.211 


736.62 


36^ 114.668 


1046.3 


30 5i 


96.604 


742.64 


36H 115.061 


1053.5 


30?^ 


96.997 


748.69 


365^ 115.454 


1060.7 


31 


97.389 


754.77 


36^ 115.846 


1068.0 


^X'A 


97.782 


760.87 


37 


116.239 


1075.2 


31K 


98.175 


766.99 


37 H 


116.632 


1082.5 


31H 


98.567 


773.14 


31 y^ 


117.024 


1089.8 


311^ 


98.960 


779.31 


31 H 


117.417 


1097.1 


31H 


99.353 


785.51 


31 Yi 


117.810 


1104.5 


SIH 


99.746 


791.73 


37 H 


118.202 


1111.8 


ZIM 


100.138 


797.98 


37 M 


118.596 


1119.2 


32 


100.531 


804.25 


3nA 


118.988 


1126.7 


32 J^ 


100.924 


810.54 


38 


119.381 


1134.1 


32 H 


101.316 


816.86 


3%H 


119.773 


1141.0 


32 H 


101.709 


823.21 


38>i 


120.106 


1149.1 


32yi 


102.102 


829.58 


38J^ 


120.559 


1156.6 


32 H 


102.494 


835.97 


38 H 


120.951 


1164.2 


32^ 


102.887 


842.39 


38H 


121.344 


1171.7 


32 H 


103.280 


848.83 


38 M 


121.737 


1179.3 


33 


103.673 


855.30 


38^ 


122.129 


1186.9 


33 H 


104.065 


861.79 


39 


122.522 


1194.6 


33 K 


104.458 


868.31 


39 H 


122.915 


1202.3 


33 H 


104.851 


874.85 


39 J^ 


123.308 


1210.0 


33 H 


105.243 


881.41 


3^H 


123.700 


1217.7 


33 H 


105.636 


888.00 


39 H 


124.093 


1225.4 


33^ 


106.029 


894.62 


39 H 


124.486 


1233.2 


33 J^ 


106.421 


901.26 


39^ 


124.878 


1241.0 


34 


106.814 


907.92 


39 J^ 


125.271 


1248.8 


34 H 


107.207 


914.61 


40 


125.664 


1256.6 


34 }i 


107.600 


921.32 


40H 


126.056 


1264.5 


34 H 


107.992 


928.06 


40 >i 


126.449 


1272.4 


34 H 


108.385 


934.82 


40 H 


126.842 


1280.3 


34 H 


108.778 


941.61 


40 H 


127.235 


1288.2 


345^ 


109.170 


948.42 


40H 


127.627 


1296.2 


34^ 


109.563 


955.25 


40 5^ 


128.020 


1304.2 


35 


106.956 


962.11 


40^ 


128.413 


1312.2 


35 H 


110.348 


969.00 


41 


128.805 


1320. a 


35 >^ 


110.741 


975.91 


41H 


129.198 


1328.3 


35 5^ 


111.134 


982.84 


41 >i 


129.591 


1336.4 


35 H 


111.527 


989.80 


41J^ 


129.983 


1344.5 


35 H 


111.919 


99rt.78 


41 H 


130.376 


1352.7 


35 M 


112.312 


1003.8 


415^ 


130.769 


1360.8 


35 3^ 


112.705 


1010.8 


41?^ 


131.161 


1369.0 


36 


113.097 


1017.9 


41H 


131.554 


1377.2 


36^ 


113.490 


1025.0 


42 


131.947 


1385.4 


36 K 


113.883 


1032.1 


42 H 


132.340 


1393.7 



STEAM ENGINES 243 

Circumferences and Area? of Circles — 'Continued 



Diam.jCircum. 


Area 


Diam. 


Circum. 


Area 


42 M 


132.732 


1402.0 


485^ 


151.189 


1819.0 


A2H 


133.125 


1410.3 


48 >i 


151.582 


1828.5 


421^ 


133.518 


1418.6 


48 H 


151.975 


1837.9 


42 s^ 


133.910 


1427.0 


48H 


152.367 


1847.5 


425^ 


134.303 


1435.4 


48 H 


152.760 


1857.0 


42 ^i 


134.696 


1443.8 


483^ 


153.153 


1866.5 


43 


135.088 


1452.2 


48 5^ 


153.545 


1876.1 


tn 


135.481 


1460.7 


49- 


153.938 


1885.7 


135.874 


1469.1 


495/8 


154.331 


1895.4 


A.ZH 


136.267 


1477.6 


49 M 


154.723 


1905.0 


43 H 


136.659 


1486.2 


49?^ 


155.116 


1914.7 


43 H 


137.052 


1494.7 


49 H 


155.509 


1924.4 


43 M 


137.445 


1503.3 


49 H 


155.902 


1934.2 


43^ 


137.837 


1511.9 


49 M 


156.294 


1943.9 


44 


138.230 


1520.5 


49 >| 


156.687 


1953.7 


44 H 


138.623 


1529.2 


50 


157.080 


1963.5 


44 >i 


139.015 


1537.9 


50 H 


157.472 


1973.3 


44^ 


139.408 


1546.6 


50 M 


157.865 


1983.2 


44 H 


139.801 


1555.3 


50 K 


158.258 


1993.1 


44 H 


140 . 194 


1564.0 


503^ 


158.650 


2003.0 


44 K 


140.586 


1572.8 


50 H 


159.043 


2012.9 


44 ?i 


140.979 


1581.6 


50 5i 


159.436 


2022.8 


45 


141.372 


1590.4 


50^ 


159.829 


2032.8 


45 H 


141.764 


1599.3 


51 


160.221 


2042.8 


45 >i 


142.157 


1608.2 


51M 


160.614 


2052.8 


45^ 


142.550 


1617.0 


51 M 


161.007 


2062.9 


45 J^ 


142.942 


1626.0 


5iy& 


161.399 


2073. a 


45 M 


143.335 


• 1634.9 


51^ 


161.792 


2083.1 


45 M 


143.728 


1643.9 


51H 


162.185 


2093. 2-. 


45^ 


144.121 


1652.9 


51^ 


162.577 


2103.3 


46 


144.513 


1661.9 


5VA 


162.970 


2113.5. 


46 H 


144.906 


1670.9 


52 


163.363 


2123.7 


46 K 


145.299 


1680.0 


52 H 


163.756 


2133.^ 


46 H 


145.691 


1689.1 


52 M 


164. 148 


2144.2 


46 H 


146.084 


1698.2 


52 H 


164.541 


2154.5 


46 H 


146.477 


1707.4 


52}^ 


164.934 


2164. S 


46 Ji 


146.869 


1716.5 


52 H 


165.326 


2175.1 


46 H 


147.262 


1725.7 


52^ 


165.719 


2185.4 


47 


147.655 


1734.9 


52^ 


166.112 


2195.8 


47 H 


148.048 


1744.2 


53 


166.504 


2206.2 


47^ 


148.440 


1753.5 


53 H 


166.897 


2216.6 


47 fi 


148.833 


1762.7 


53 >i 


167.290 


2227.0 


47 >^ 


149.226 


1772.1 


53^ 


167.683 


2237.5 


47 H 


149.618 


1781.4 


53 H 


168.075 


2248.0 


47% 


150.011 


1790.8 


53H 


168.468 


2258.5 


47 Ji 


150.404 


1800.1 


53% 


168.861 


2269.1 


48 


150.796 


1809.6 


53?^ 


169.253 2279.6 



244 STEAM ENGINEERING 

Circumferences and Areas of Circles — Continued 



Diam. 


Circum. 


Area 


Diam. 


Circum. 


Area 


54 


169.646 


2290.2 


59^ 


187.710 


2803.9 


54H 


170.039 


2300.8 


59;^ 


188.103 


2815.7 


54 M 


170,431 


2311.5 


60 


188.496 


2827.4 


54 H 


170.824 


2322.1 


60H 


188.888 


2839.2 


54 H 


171.217 


2332.8 


&0}4 


189.281 


2851.0 


54 H 


171.609 


2343.5 


QOH 


189.674 


2862.9 


54 M 


172.002 


2354.3 


60 H 


190.066 


2874.8 


54: pi 


172.395 


2365.0 


60 H 


190.459 


2886.6 


55 


172.788 


2375.8 


60^ 


190.852 


2898.6 


55 K 


173.180 


2386.6 


60 2^ 


191.244 


2910.5 


55 J^ 


173.573 


2397.5 


61 


191.637 


2922.5 


55 H 


173.966 


2408.3 


Qiys 


192.030 


2934.5 


55 M 


174.358 


2419.2 


61M 


192.423 


2946.5 


55 H 


174.751 


2430.1 


61 H 


192.815 


2958.5 


55 3^ 


175.144 


2441.1 


61M 


193.208 


2970.6 


55?^ 


175.536 


2452.0 


61 H 


193.601 


2982.7 


56 


175.929 


2463.0 


61 M 


193.993 


2994.8 


56 H 


176.322 


2474.0 


6114 


194.386 


3006.9 


56 >i 


176.715 


2485.0 


62 


194.779 


3019.1 


561^ 


177.107 


2496.1 


62 M 


195.171 


3031.3 


56 K 


177.500 


2507.2 


62 M 


195.564 


3043.5 


56 H 


177.893 


2518.3 


62 H 


195.957 


3055.7 


56 M 


178.285 


2529.4 


62 M 


196.350 


3068.0 


56 2^ 


178.678 


2540.6 


62^ 


196.742 


3080.3 


57 


179.071 


2551.8 


62 M 


197.135 


3092.6 


57 H 


179.463 


2563.0 


62 J^ 


197.528 


3104.9 


57 >i 


179.856 


2574.2 


63 


197.920 


3117.2 


57 fi 


180.249 


2585.4 


63^ 


198.313 


3129.6 


57 H 


180.642 


2596.7 


63 >i 


198.706 


3142.0 


57 ys 


181.034 


2608.0 


63 H 


199.098 


3154.5 


57 H 


181.427 


2619.4 


62, Vi 


199.491 


3166.9 


57 H 


181.820 


2630.7 


63 H 


199.884 


3179.4 


58 


182.212 


2642.1 


63 M 


200.277 


3191.9 


58 >i 


182.605 


2653.5 


63^8 


200.669 


3204.4 


58 M 


182.998 


2664.9 


64 


201.062 


3217.0 


58 H 


183.390 


2676.4 


64^ 


201.455 


3229.6 


58 V^ 


183.783 


2687.8 


64 M 


201.847 


3242.2 


58 H 


184.176 


2699.3 


64 H 


202.240 


3254.8 


585^ 


184.569 


2710.9 


64 J^ 


202.633 


3267.5 


58 J^ 


184.961 


2722.4 


64S/i 


203.025 


3280.1 


59 


185.354 


2734.0 


64 M 


203.418 


3292.8 


595^ 


185.747 


2745.6 


64H 


203.811 


3305.6 


59 K 


186.139 


2757.2 


65 


204.204 


3318.3 


59^ 


186.532 


2768.8 


65 H 


204.596 


3331.1 


591^ 


186.925 


2780.5 


65 H 


204.989 


3343.9 


59 H 


187.317 


2792.2 


6b H 


205.382 


3356.7 



STEAM ENGINES 245 

Circumferences and Areas of Circles — Continued 



Diam. 


Circum. 


Area 


Diam. 


Circum, 


Area 


65 H 


205.774 


3369.6 


71 H 


224.231 


4001.1 


65 H 


206.167 


3382.4 


71H 


224.624 


4015.2 


65 M 


206.560 


3395.3 


71^ 


225.017 


4029.2 


65^8 


206.952 


3408.2 


71H 


225.409 


4043.3 


66 


207.345 


3421.2 


71V& 


225.802 


4057.4 


66M 


207.738 


3434.2 


72 


226.195 


4071.5 


66 H 


208.131 


3447.2 


72 H 


226.587 


4085.7 


66 H 


208.523 


3460.2 


72 M 


226.980 


4099.8 


66 Ji 


208.916 


3473.2 


72 H 


227.373 


4114.0 


66^ 


209.309 


3486.3 


72 >^ 


227.765 


4128.2 


66 M 


209.701 


3499.4 


72 H 


228.158 


4142.5 


66?^ 


210.094 


3512.5 


72 M 


228.551 


4156.8 


67 


210.487 


3525.7 


72^ 


228.944 


4171.1 


67 H 


210.879 


3538.8 


73 


229.336 


4185.4 


67 M 


211.272 


3552.0 


73^ 


229.729 


4199.7 


67 H 


211.665 


3565.2 


73 >^ 


230.122 


4214.1 


673^ 


212.058 


3578.5 


73 H 


230.514 


4228.5 


67 H 


212.450 


3591.7 


73 J^ 


230.907 


4242.9 


67 % 


212.843 


3605.0 


73 H 


231.300 


4257.4 


67^ 


213.236 


3618.3 


73 M 


231.692 


4271.8 


68 


213.628 


3631.7 


73^ 


232.085 


4286.3 


68 H 


214.021 


3645.0 


74 


232.478 


4300.8 


68 M 


214.414 


3658.4 


74 Vs 


232.871 


4315.4 


68 H 


214.806 


3671.8 


74 3I 


233.263 


4329.9 


683^ 


215.199 


3685.3 


74 H 


233.656 


4344.5 


68 H 


215.592 


3698.7 


741^ 


234.049 


4359.2 


68 34 


215.984 


3712.2 


74 H 


234.441 


4373.8 


683^8 


216.377 


3725.7 


7434 


234.834 


4388.5 


69 


216.770 


3739.3 


74?^ 


235.227 


4403.1 


69 K 


217.163 


3752.8 


75 


235.619 


4417.9 


69 M 


217.555 


3766.4 


75^8 
75 M 


236.012 


4432.6 


69^ 


217.948 


3780.0 


236.405 


4447.4 


691^ 


218.341 


3793.7 


75 H 


236.798 


4462.2 


69 H 


218.733 


3807.3 


751^ 


237.190 


4477.0 


69 M 


219.126 


3821.0 


75 H 


237.583 


4491.8 


69^ 


219.519 


3834.7 


75 M 


237.976 


4506.7 


70 


219.911 


3848.5 


7bVi 


238.368 


4521.5 


70 H 


220.304 


3862.2 


76 


238.761 


4536.5 


70^ 


220.697 


3876.0 


76% 


239.154 


4551.4 


70^ 


221.090 


3889.8 


76 M 


239.546 


4566.4 


70 J^ 


221.482 


3903.6 


76 H 


239.939 


4581.3 


70 H 


221.875 


3917.5 


76^ 


240.332 


4596.3 


70^ 


222.268 


3931.4 


76 H 


240.725 


4611.4 


70^ 


222.660 


3945.3 


763^ 


241.117 


4626.4 


71 


223.053 


3959.2 


76?^ 


241.510 


4641.5 


71^ 


223.446 


3973.1 


77 


241.903 


4656.6 


71M 


223.838 


3987.1 


77% 


242.295 


4671.8 



246 STEAM ENGINEERING 

Circumferences and Areas of Circles — Continued ' 



Diam. 


Circum. 


Area 


Diam. 


Circum. 


Area 


77 K 


242.688 


4686.9 


83 Vs 


261.145 


5426.9 


77 H 


243.081 


4702.1 


83 M 


261.538 


5443.3 


77 Yi 


243.473 


4717.3 


83 H 


261.930 


5459.6 


77 H 


243.866 


4732.5 


83 H 


262.323 


5476.0 


77 M 


244.259 


4747.8 


83^8 


262.716 


5492.4 


77^ 


244.652 


4763 . 1 


83 H 


263.108 


5508.8 


•78 


245.044 


4778.4 


83 ys 


263.501 


5525.3 


78% 


245.437 


4793.7 


84 


263.894 


5541.8 


783^ 


245.830 


4809.0 


84 Vs 


264.286 


5558.3 


783/i 


246.222 


4824.4 


84: H 


264.679 


5574.8 


783^ 


246.615 


4839.8 


84: H 


265.072 


5591.4 


78^ 


247.008 


4855.2 


84y2 


265.465 


5607.9 


78 M 


247.400 


4870.7 


84 H 


265.857 


5624.5 


78 H 


247.793 


4886.2 


84 M 


266 . 250 


5641.2 


79 


248.186 


4901.7 


84 J^ 


266.643 


5657.8 


79 Vs 
79 M 


248.579 


4917.2 


85 


267.035 


5674.5 


248.971 


4932.7 


85 Vs 


267.428 


5691.2 


79 H 


249.364 


4948.3 


85 S 


267.821 


5707.9 


79 H 


249.757 


4963.9 


85^ 


268.213 


5724.7 


79 H 


250 . 149 


4979.5 


85 H 


268.606 


5741.5 


79 M 


250.542 


4995.2 


85 s^ 


268.999 


5758.3 


79 >i 


250.935 


5010.9 


85 3^ 


269.392 


5775.1 


80 


251.327 


5026.5 


85 >g 


269.784 


5791.9 


80 Vs 


251.720 


5042.3 


86 


270.177 


5808.8 


80K 


252.113 


5058.0 


86% 


270.570 


5825.7 


80H 


252.506 


5073.8 


86 M 


270 . 962 


5842.6 


80 >^ 


252.898 


5089.6 


86 H 


271.355 


5859.6 


80 H 


253.291 


5105.4 


86 Vi 


271.746 


5876.5 


«0M 


253.684 


5121.2 


86 H 


272.140 


5893.5 


80^ 


254.076 


5137.1 


86^ 


272.533 


5910.6 


81 


254.469 


5153.0 


86^ 


272.926 


5927.6 


81% 


254.862 


5168.9 


87 


273.319 


5944.7 


81^ 


255.254 


5184.9 


87 Vs 


273.711 


5961.8 


81 H 


255.647 


5200.8 


87 M 


274.104 


5978.9 


81^ 


256.040 


5216.8 


87 H 


274.497 


5996.0 


81 H 


256.433 


5232.8 


87 Yi 


274 . 889 


6013.2 


81 M 


256.825 


5248.9 


87 H 


275.282 


6030.4 


81J^ 


257.218 


5264.9 


87 3^ 


275 . 675 


6047.6 


.82 


257.611 


5281.0 


87?/8 


276.067 


6064.9 


82 Vs 


258.003 


5297 . 1 


88 


276.460 


6082.1 


82 M 


258.396 


5313.3 


88% 


276.853 


6099.4 


82^ 


258.789 


5329.4 


88 >| 


277.246 


6116.7 


821^ 


259.181 


5345 . 6 


88?^ 


277.638 


6134.1 


82 H 


259.574 


5361.8 


881^ 


278.031 


6151.4 


82 M 


259.967 


5378.1 


88 H 


278.424 


6168.8 


82^ 


260 . 359 


5394.3 


88 M 


278.618 


6186.2 


83 


260.752 


5410.6 


88 >i 


279.209 


6203.7 



STEAM ENGINES 247 

,Cii6umferences and Areas of Circles — Continued 



Diam. 


Circum. 


Area 


Diam. 


Circum. 


Area 


89 


279.602 


6221.1 


94 H 


298.059 


7069.6 


89% 
89 k 


279.994 


6238.6 


95 


298.451 


7088.2 


280.387 


6256.1 


95% 


298.844 


7106.9 


89 H 


280.780 


6273.7 


95% 


299.237 


7125.6 


89 H 

89 k 


281.173 


6291.2 


95 H 


299 . 629 


7144.3 


281.565 


6308.8 


95y2 


300.022 


7163.0 


89 M 


281.958 


6326.4 


9b SA 


300.415 


7181.8 


89?^ 


282.351 


6344.1 


95% 


300.807 


7200.6 


60 


282.743 


6361.7 


95^ 


301.200 


7219.4 


^^ 


283.136 


6379.4 


96 


301.593 


7238.2 


283.529 


6397.1 


96% 


301.986 


7257 . 1 


90 ys 


283.921 


.6414.9 


96% 


302.378 


7276.0 


90H 


284.314 


6432.6 


96^ 


302.771 


7294.9 


90 H 


284.707 


6450.4 


961^ 


303.164 


7313.8 


9^M 


285 . 100 


6468.2 


96^ 


303.556 


7332.8 


285.492 


6486.0 


96% 


303.949 


7351.8 


91 


285.885 


6503.9 


96^ 


304.342 


7370.8 


tm 


286.278 


6521.8 


97 


304.734 


7389.8 


286.670 


6539.7 


97% 


305.127 


7408.9 


91H 


287.063 


6557 . 6 


97% 


305.520 


7428.0 


91 H 


287.456 


6575.5 


97 H 


305.913 


7447.1 


91H 


287.848 


6593.5 


97 H 


306.305 


7466.2 


91% 


288.241 


6611.5 


91 H 


306.698 


7485.3 


91?^ 


288.634 


6629.6 


97% 


307.091 


7504.5 


92- 


289.027 


6647.6 


97 Vi 


307.483 


7523.7 


i^ 


289.419 


6665.7 


98 - 


307.876 


7543.0 


289.812 


6683.8 


98% 


308.269 


7562.2 


92 H 


290.205 


6701.9 


98% 


308.661 


7581.5 


921^ 


290.597 


6720.1 


98 H 


309.054 


7600.8 


92 H 


290.990 


6738.2 


9sy2 


309.447 


7620.1 


928^ 


291.383 


6756.4 


98 H 


309.840 


7639.5 


92?^ 


291.775 


6774.7 


98% 


310.232 


7658.9- 


93 


292.168 


6792.9 


98^4 


310.625 


7678.3 


93% 


292.561 


6811.2 


99 


311.018 


7697.7 


93)^ 


292.954 


6829.5 


99% 


311.410 


7717.1 


93 H 


293.346 


6847.8 


99% 


311.803 


7736.6 


931^ 


293.739 


6866.1 


99^ 


312.196 


7756.1 


93 H 


294.132 


6884.5 


99 H 


312.588 


7775.6 


93% 


294.524 


6902.9 


99 H 


312.981 


7795.2 


93^ 


294.917 


6921.3 


99 H 


313.374 


7814.8 


94 


295.310 


6939.8 


99 ys 


313.767 


7834.4 


MV& 


295.702 


6958.2 


100 


314.159 


7854.0 


94: H 


296.095 


6976.7 








94^ 


296.488 


6995.3 








nn 


296.881 


7013.8 








297.273 


7032.4 








94% 


297.666 


705 rro 









248 STEAM ENGINEERING 

DETERMINING THE MEAN EFFECTIVE 
PRESSURE 

There are many cases in which it be- 
comes necessary to determine the mean 
effective pressure that will result from 
given conditions in advance of the in- 
stallation of the engine, or in other words 
before an indicator diagram can be se- 
cured. As this is an important matter 
it will be explained in detail. 

The terms "average pressure" and 
"mean effective pressure" are used 
indiscriminately, not only by engineers 
in daily practice, but in mechanical 
books and papers by men who in some 
cases do not understand the difference, 
and in others are too careless to properly 
separate them. The effect in either case 
on the student is precisely the same, as 
it gives him a wrong idea of the whole 
subject. 

The average pressure acting on the 
piston of a steam engine is found by 
taking into consideration the initial 
pressure, or pressure at the beginning 
of the stroke in connection with the pres- 
sure realized at short succeeding inter- 
vals until the end is reached. When 
the average of them is found it consti- 
tutes the average pressure for those 
conditions, and this must necessarily be 
the pressure above a perfect vacuum in 
all cases. 



STEAM ENGINES 



249 



Special attention is called to the fact 
that the back pressure is not considered, 
as it can have no effect on the average 




<}> 



pressure, because they act on opposite 
sides of the piston. 

Fig. 38 is a theoretical indicator dia- 
gram in which the full lines only are used 
to determine the average pressure. The 



250 STEAM ENGINEERING 

dotted lines are inserted to make the 
diagram complete, but they have no 
other use here. This 'diagram was orig- 
inally laid out accurately by a No. 40 
scale, showing an initial pressure of 120 
pounds absolute with the point of cut-off 
at one-quarter stroke. While it is nec- 
essary to reduce the size of this diagram 
for use here, the correct proportions are 
preserved, but it is not practicable to 
measure it with a No. 40 scale now. 
When the ordinates are laid out as shown 
their total length in inches ascertained 
and divided by the number used, the 
quotient is the average height of the 
diagram. Multiply this by the scale 
adopted (which is No, 40 in this case), 
and the product is the average pressure 
acting on the engine piston. These 
ordinates are the exact length to be 
measured in order to make the matter 
as plain as possible. The atmospheric 
line is shown at AA and the perfect 
vacuimi line at VV. 

While a diagram clearly illustrates the 
principle involved, the same result can 
be secured with less trouble by calcula- 
tion, using the following rule: 

Ascertain the ratio of expansion and 
find the corresponding hyperbolic logar- 
ithm in a table prepared for this pur- 
pose. Add 1, multiply the sum by the ini- 
tial pressure absolute, and divide by the 



STEAM ENGINES 251 

ratio of expansion. The quotient is the 
average pressure for stated conditions. 

When given as a formula it appears as 
follows: 

(HypLog+l)XP_. 
R -^- 

Hyp Log = Hyperbolic logarithm of 
the ratio of expansion. 
P = Initial pressure absolute. 
R= Ratio of expansion. 
A = Average pressure. 
For illustration and explanation take 
a plant carrying 110 pounds pressure by 
the gauge, or 125 pounds absolute, with 
an initial pressure of 120 pounds in the 
cylinder, cutting off steam at one-quar- 
ter or .25 of the stroke. 

Under these conditions the ratio of 
expansion is 4 the hyperbolic logarithm 
of which is 1.3863 and when due appli- 
cation is made of the formula it gives 
the following result: 

(1.3863+1) X120 ^^ ^^ 
4 
pounds average pressure. 

Fig. 39 illustrates the proper method 
for finding the mean effective pressure 
from an indicator diagram. It is a 
reproduction of Fig. 38, except that 
there are no dotted lines, and the ordi- 
nates are shortened to show the exact 



252 



STEAM ENGINEERING 



measurements required When the 
total length of these ordinates in inches 
is found and divided by the number 




used, the quotient is the mean effective 
height of the diagram, and when this is 
multiplied by the scale used, the pro- 
duct is the mean effective pressure. 
As these ordinates are shorter than 



STEAM ENGINES 253 

before, the effect is to cat off or subtract 
the back pressure. No stated number 
of ordinates is required, as the desired 
result is found by dividing the total 
length by the number adopted. 

Examination of the two preceding 
diagrams illustrates the fact that when 
the rule for finding the average pressure 
on the piston of an engine is stated, it is 
only necessary to add another clause and 
it becomes a rule for calculating the 
mean effective pressure as follows: 

Ascertain the ratio of expansion and 
find the corresponding hyperbolic logar- 
ithm in a table prepared for this purpose. 
Add 1, multiply the sum by the initial 
pressure absolute, divide by the ratio of 
expansion, and subtract the back pres- 
sure. The remainder is the mean effec- 
tive pressure for stated conditions. 

When given as a formula it appears 
as follows: 

(HypLog+l)XP _^^^^p 

Hyp Log = Hyperbolic logarithm of 

the ratio of expansion. 

P = Initial pressure absolute. 

R = Ratio of expansion. 

B = Back pressure above a 

vacuum. 

M E P = Mean effective pressure 

for given conditions, 



254 STEAM ENGINEERING 

Applying this formula to a case where 
the initial pressure is 120 pounds abso- 
lute, the ratio of expansion 4 and the 
back pressure 16 pounds gives the fol- 
lowing result: 

(1.3863 + l)Xl20_^g^^^^^g 



4 
pounds mean effective pressure. 

When calculating the size of an engine 
that would be required to develop 300 
horse power, the mean effective pressure 
was assumed to be 50 pounds. As the 
above result is more than the required 
pressure, the point of cut off would be a 
trifle shorter than at one- quarter stroke, 
which is satisfactory, as it will give a 
lower terminal pressure. 

THE RATIO OF EXPANSION AND THE BACK 
PRESSURE 

When the cut-off valve on an auto- 
matic, or any other kind of an engine, 
closes, the piston has travelled a certain 
part of the stroke, and the relation that 
this part bears to the whole stroke is the 
ratio of expansion. For illustration, 
suppose that the cut-off valve closes 
when the piston has travelled one-quar- 
ter or .25 of the stroke. Taking the 
whole stroke as unity or 1 and dividing 



STEAM ENGINES 255 

it by .25 gives the ratio of expansion. 
Then 1-=-. 25 = 4. 

For estimating the mean effective 
pressure the clearance is not taken into 
account, as it makes the calculation 
more complicated for what is not in- 
tended to be an accurate computation. 
If greate accuracy is required, the clear- 
ance should be stated as the percentage of 
the whole stroke. It is then added to 
unity or 1 which represents the whole 
stroke, also to the fraction which 
represents the point of cut off and 
the former is divided by the latter 
as before. 

Suppose that the clearance is 5 per 
cent, or .05 of the whole stroke, and the 
cut off takes place at one-quarter stroke. 
The ratio of expansion is then: 



l + -°^- = 3.5 



.25 + .05 

The next table gives the actual ratios 
of expansion taking clearance from 1 to 
10 per cent, into consideration. The 
first column gives the point of cut off as 
a fraction of the whole stroke. The 
second gives the ratio of expansion with- 
out clearance, and succeeding columns 
give the ratio with the amount of clear- 
ance given for each as a per cent, of the 
whole stroke. 



256 


STEAM ENGINEERING 


J 


ACTUAL RATIOS OF EXPANSION. ij 




Per cent, of Clearance. 


:i 


PofO 





1 


2 


3 


4 ': 


.01 


100.00 


50.5 


34.0 


25.75 


20.8 


.02 


50.00 


33.67 


25.50 


20.60 


17.33, 


.03 


33.33 


25.25 


20.40 


17.16 


14.86 


.04 


25.00 


20.20 


17.00 


14.71 


13.00 


.05 


20.00 


16.83 


14.57 


12.87 


11.55 


.06 


16.67 


14.43 


.12.75 


11.44 


10.40 


.07 


14.28 


12.62 


11.33 


10.30 


9.46 


.08 


12.60 


11.22 


10.2 


9.36 


8.67 


.09 


11.11 


10.10 


9.27 


8.58 


8.00 


.10 


10.00 


9.18 


8.50 


7.92 


7.43 


.11 


9.09 


8.42 


7.84 


7.36 


6.93 


.12 


8.33 


7.78 


7.25 


6.86 


6.50 


.14 


7.14 


6.73 


6.37 


6.06 


5.78 


.16 


6.25 


5.94 


5.67 


5.42 


5.20 


.20 


5.00 


4.81 


4.64 


4.48 


4.33 


.25 


4.00 


3.88 


3.77 


3.68 


3.58 


.30 


3.33 


3.26 


3.19 


3.12 


3.06 


.40. 


2.50 


2.46 


2.43 


2.40 


2.36 


.50 


2.00 


1.98 


1.96 


1.94 


1.92 


.60 


1.67 


1.66 


1.65 


1.64 


1.63 


.70 


1;43 


1.42 


1.42 


1.41 


1.41 


.80 


1.25 


1.25 


1.244 


1.241 


1.238 


.90 


1.111 


1.11 


1.109 


1.108 


1.106 


1.00 


1.00 


1.00 


1,000 


1.000 


1.000 



ACTUAL RATIOS OF EXPANSION. 
Per cent, of 'Clearance. 



Pof C 


5 


6 


7 


8 


9 


10 


.01 


17.5 |15.14 


13.38 


12.00 


'0.9 


10. 


.02 


15.00 1 13.25 


11.89 


10.80 


9.91 


8.17 


.03 


13.12 11.78 


]0.70 


9.82 


9.08 


8.46 


.04 


11.66 10.60 


9.73 


9.00 


8.39 


7.86 


.05 


10.50 9.64 


8.92 


8.31 


7.79 


7.33 


.06 


9.55 


8.33 


8.23 


7.71 


7.27 


6.88 


.07 


8.75 


8.15 


7.64 


7.20 


6.81 


6.47 


.08 


8.08 


7.57 


7.13 


6.75 


6.41 


6.11 


.09 


7.50 


7.07 


6.69 


6.35 


6.06 


5.79 


.10 


7.00 


6.62 


6.30 


6.00 


5.74 


5.50 


.11 


6.56 


6.24 


5.94 


5.68 


5.45 


5.24 



STEAM ENGINES 



257 



Actual Ratios of Expansion — Continued 
Per cent, of Clearance 



PofC 


5 


6 


7 


8 


9 


10 


.12 


6.18 


5.89 


5.63 


5.40 


5.19 


5.00 


.14 


5.53 


5.30 


5.10 


4.91 


4.74 


4.58 


.16 


5.00 


4.82 


4.65 


4.50 


4.36 


4.23 


.20 


4.20 


4.08 


3.96 


3.86 


3.76 


3.61 


.25 


3.50 


3.42 


3.34 


3.27 


3.21 


3.14 


.30 


3.00 


2.94 


2.90 


2.84 


2.80 


2 75 


.40 


2.33 


2.30 


2.28 


2.25 


2.22 


2^20 


.50 


1.90 


1.89 


1.88 


1.86 


1.85 


1.83 


.60 


1.615 


1.606 


1.597 


1.588 


1.580 


1.571 


.70 


1.400 


1.395 


1.390 


1.385 


1.380 


1.375 


.80 


1.235 


1.233 


1.230 


1.227 


1.224 


1.222 


.90 


1.105 


1.104 


1.103 


1.102 


1.101 


1.100 


1.00 


1.000 


1.000 


1.000 


1.000 


1.000 


1.000 



Having determined the assumed or 
the actual ratio of expansion, it becomes 
necessary to find the corresponding 
hyperbolic logarithm in a table, and add 
one to it in each case. The reason for 
this is because the work done by the 
steam from the beginning of the stroke 
to the point of cut off is represented by 
unity or 1 and the work performed after 
the cut off has taken place, or during 
expansion of the steam, is shown com- 
paratively by the hyperbolic logarithm 
of the ratio of expansion. As these 
must be added in order to get the total 
work done during the whole stroke it 
appears in the rules and formulas given , 
to make the process complete. 

This is illustrated in Fig. 40 which 
shows a -cut off at one-quarter stroke 



258 



STEAM ENGINEERING 



which makes the expansion rate 4 as 
previously explained. The vertical line 
extends from the point of cut off to the 




vacuum line, and the space to the right 
of it represents work done by steam 
direct from the boiler, and this is unity 
or 1. The space to the left shows work ^ 



b 



STEAM ENGINES 259 

done by expansion of the steam, and it is 
represented by the hyperbolic logarithm 
of 4, which is 1.3863. When these two 
factors are added the sum is the total 
work done during the complete stroke. 
These measurements are taken to the 
vacuum line, because it is necessary in 
order to include the whole. The counter 
pressure line and the atmospheric line 
are shown in dotted form to make the 
diagram complete, but their only prac- 
tical use in this connection is to show 
how the total work is divided, as they 
cannot make it either more or less. 

Where an engine is located at or near 
the sea level, and is run non-condensing, 
the back pressure should be estimated 
at not less than 16 pounds, as the atmos- 
phere causes nearly 15 and the remain- 
der is allowed for friction in exhaust 
passages and pipes, as the steam travels 
to the outer air. If a condenser is to be 
used under good conditions, the back 
pressure may be taken at 3 pounds, as 
that represents good practice. 

The following table contains hyper- 
bolic logarithms of numbers from 1.01 
to 20 to be used in connection with rules 
and formulas for determining the aver- 
age and the mean effective pressures. 
When the ratio of expansion is deter- 
mined, find the nxmiber in this table and 
the corresponding hyperbolic logarithm 



260 


STEAM ENGINEERING 




will be found 


in the next column, and 


this is 


to be used as 


directed. 


; 


HYPERBOLIC LOGARITHMS OF NUMBERS | 




from 1.01 to 20.00. 




No.- 


Log 


No. 


Log. 


No. 


Log. 


1.01 


.0099 


1.43 


.3577 


1.85 


.6152 


1.02 


.0198 


1.44 


.3646 


1.86 


.6206 


1.03 


.0296 


1.45 


.3716 


1.87 


.6259 


1.04 


.0392 


1.46 


.3784 


1,88 


.6313 


1.05 


.0488 


1.47 


.3853 


1.89 


.6336 


1.06 


.0583 


1.48 


.3920 


1.90 


.6419 


1.07 


.0677 


1.49 


.3988 


1.91 


.6471 


1.08 


.0770 


1.50 


.4055 


1.92 


.6523 ! 


1.09 


.0862 


1.51 


.4121 


1.93 


.6575 i 


1.10 


.0953 


1.52 


.4187 


1.94 


.6627 


1.11 


.1044 


1.53 


.4253 


1.95 


.6678 


1.12 


.1133 


1.54 


.4318 


1.96 


.6729 


1.13 


.1222 


1.65 


.4383 


1.97 


.6780 ! 


1.14 


.1310 


1.56 


.4447 


1.98 


.6831 ! 
.6881 


1.15 


.1398 


1.57 


.4511 


1.99 


1.16 


.1484 


1.58 


.4574 


2.00 


.6931 


1.17 


.1570 


1.59 


.4637 


2.01 


.6981 


1.18 


.1655 


1.60 


.4700 


2.02 


.7031 


1.19 


.1740 


1.61 


.4762 


2.03 


.7080- 


1.20 


.1823 


1.62 


.4824 


2.04 


.7129 


1.21 


.1906 


1.63 


.4886 


2.05 


.7178 ' 


1.22 


.1988 


1.64 


.4947 


2.06 


.7227 


1.23 


.2070 


1.65 


.5008 


2.07 


.7275 


1.24 


.2151 


1.66 


.5068 


2.08 


.7324 


1.25 


.2231 


1.67 


.5128 


2.09 


.7372 


1.26 


.2311 


1.68 


.5188 


2.10 


.7419 


1.27 


.2390 


1.69 


.5247 


2.11 


.7467 


1.28 


.2469 


1.70 


.5306 


2.12 


.7514 


1.29 


.2546 


1.71 


.5365 


2.13 


.7561 


1.30 


.2624 


1.72 


.5423 


2.14 


.7608 


1.31 


.2700 


1.73 


.5481 


2.15 


.7655 


1.32 


.2776 


1.74 


.5539 


2.16 


.7701 


1.33 


.2852 


1.75 


.5596 


2.17 


.7747 


1.34 


.2927 


1.76 


.5653 


2.18 


.7793 


1.35 


.3001 


1.77 


5710 


2.19 


.7839 


1.36 


.3075 


1.78 


•5766 


2.20 


.7885 


1.37 


.3148 


1.79 


•5822 


2.21 


.7930 


1.38 


.3221 


1.80 


•5878 


2.22 


.7975 


1.39 


.3293 


1.81 


•5933 


2.23 


.8020 


1.40 


.3365 


1.82 


•5988 


2.24 


.8065 


1.41 


.3436 


1.83 


•6043 


2.25 


.8109 


1.42 


.3507 


1.84 


•6098 


2.26 


.8154 



STEAM ENGINES 



261 



Hyperbolic Logarithms of Numbers — Continued 



No. 


Log. 


No.^ 


Log. 


No. 


Log. 


2.27 


.8198 


2.74 


1.0080 


3.21 


1.1663 


2.28 


.8242 


2.75 


1.0116 


3.22 


1.1694 


2.29 


.8286 


2.76 


1.0152 


3.23 


1.1725 


2.30 


.8329 


2.77 


1.0188 


3.24 


1.1756 


2.31 


.8372 


2.78 


1.0225 


3.25 


1.1787 


2.32 


.8416 


2.79 


1.0260 


3.26 


1.1817 


2.33 


.8458 


2.80 


1.0296 


3.27 


1 . 1848 


2.34 


.8502 


2.81 


1.0332 


3.28 


1.1878 


2.35 


.8544 


2.82 


1.0367 


3.29 


1.1909 


2.36 


.8587 


2.83 


1.0403 


3.30 


1.1939 


2.37 


.8629 


2.84 


1.0438 


3.31 


1.1969 


2.38 


.8671 


2.85 


1.0473 


3.32 


1.1999 


2.39 


.8713 


2.86 


1.0508 


3.33 


1.2030 


2.40 


.8755 


2.87 


1.0543 


3.34 


1.2060 


2.41 


.8796 


2.88 


1.0578 


3.35 


1.2090 


2.42 


.8838 


2.89 


1.0613 


3.36 


1.2119 


2.43 


.8879 


2.90 


1.0647 


3.37 


1.2149 


2.44 


.8920 


2.91 


1.0682 


3.38 


1.2179 


2.45 


.8961 


2.92 


1.0716 


3.39 


1.2208 


2.46 


.9002 


2.93 


1.0750 


3.40 


1.2238 


2.47 


.9042 


2.94 


1.0784 


3.41 


1.2267 


2.48 


.9083 


2.95 


1.0813 


3.42 


1.2296 


2.49 


.9123 


.2,96 


1.0852 


3.43 


1.2326 


2.50 


.9163 


2.97 


1.0886 


3.44 


1.2355 


2.51 


.9203 


2.98 


1.0919 


3.45 


1.2384 


2.52 


.9243 


2.99 


1.0953 


3.46 


1.2413 


2.53 


.9282 


3.00 


1.0986 


3.47 


1.2442 


2.54 


.9322 


3.01 


1.0009 


3.48 


1.2470 


2.55 


.9361 


3.02 


.1.1053 


3.49 


1.2499 


2.56 


.9400 


3.03 


1 . 1086 


3.50 


1.2528 


2.57 


.9439 


3.04 


1.1119 


3.51 


1.2556 


2.58 


.9478 


3.05 


1.1151 


3.52 


1.2585 


2.59 


.9517 


3.06 


1.1184 


3.53 


1.2613 


2.60 


.9555 


3.07 


1.1217 


3.54 


1.2641 


2.61 


.9594 


3.08 


1.1249 


3.55 


1.2669 


2.62 


.9632 


3.09 


1.1282 


3.56 


1.2698 


2.63 


.9670 


3.10 


1.1314 


3.57 


1.2726 


2.64 


.9708 


3.11 


1 . 1346 


3.58 


1.2754 


2.65. 


.9746 


3.12 


1.1378 


3.59 


1.2782 


2.66 


.9783 


3.13 


1.1410 


3.60 


1.2809 


2.67 


.9821 


3.14 


1.1442 


3.61 


1.2837 


2.68 


.9858 


3.15 


1.1474 


3.62 


1.2865 


2.69 


.9895 


3.16 


1.1506 


3.63 


1 . 2892 


2.70 


.9933 


3.17 


1.1537 


3.64 


1.2920 


2.71 


.9969 


3.18 


1 . 1569 


3.65 


1.2947 


2.72 


1.0006 


3.19 


1.1600 


3.66 


1.2975 


2.73 


1.0043 


3.20 


1.1632 


3.67 


1.3002 



262 STEAM ENGINEERING 

Hyperbolic Logarithms of Numbers — Continued 



No. 


Log. 


No. 


Log. 


No. 


Log. 


3.68 


1.3029 


4.13 


1.4183 


4.58 


1.5217 


3.69 


1.3056 


4.14 


1.4207 


4.59 


1.5239 


3.70 


1.3083 


4.15 


1.4231 


4.60 


1.5261 


3.71 


1.3110 


4.16 


1.4255 


4.61 


1.5282 


3.72 


1.3137 


4.17 


1.4279 


4.62 


1.5304 


3.73 


1.3164 


4.18 


1.4303 


4.63 


1.5326 


3.74 


1.3191 


4.19 


1.4327 


4.64 


1.5347 


3.75 


1.3218 


4.20 


1.4351 


4.65 


1.5369 


3.76 


1.3244 


4.21 


1.4375 


4.66 


1.5390 


3.77 


1.3271 


4.22 


1.4398 


4.67 


1.5412 


3.78 


1.3297 


4.23 


1.4422 


4.68 


1.5433 


3.79 


1.3324 


4.24 


1.4446 


4.69 


1.5454 


3.80 


1.3350 


4.25 


1.4469 


4.70 


1.5476 


3.81 


1.3376 


4.26 


1.4493 


4.71 


1.5497 


3.82 


1.3403 


4.27 


1.4516 


4.72 


1.5518 


3.83 


1.3429 


4.28 


1.4540 


4.73 


1.5539 


3.84 


1.3455 


4.29 


1.4563 


4.74 


1.5560 


3.85 


1.3481 


4.30 


1.4586 


4.75 


1.5581 


3.86 


1.3507 


4.31 


1.4609 


4.76 


1.5602 


3.87 


1.3533 


4.32 


1.4633 


4.77 


1.5623 


3.88 


1.3558 


4.33 


1.4656 


4.78 


1.5644 


3.89 


1.3584 


4.34 


1.4679 


4.79 


1.5665 


3.90 


1.3610 


4.35 


1.4702 


4.80 


1.5686 


3.91 


1.3635 


4.36 


1.4725 


4.81 


1.5707 


3.92 


1.3661 


4.37 


1.4748 


4.82 


1.5728 


3.93 


1.3686 


4.38 


1.4770 


4.83 


1.5748 


3.94 


1.3712 


4.39 


1.4793 


4.84 


1.5769 


3.95 


1.3737 


4.40 


1.4816 


4.85 


1.5790 


3.96 


1.3762 


4.41 


1.4839 


4.86 


1.5810 


3.97 


1.3788 


4.42 


1.4861 


4.87 


1.5831 


3.98 


1.3813 


4.43 


1.4884 


4.88 


1.5851 


3.99 


1.3838 


4.44 


1.4907 


4.89 


1.5872 


4.00 


1.3863 


4.45 


1.4929 


4.90 


1.5892 


4.01 


1.3888 


4.46 


1.4951 


4.91 


1.5913 


4.02 


1.3913 


4.47 


1.4974 


4.92 


1.5933 


4.03 


1.3938 


4.48 


1.4996 


4.93 


1.5953 


4.04 


1.3962 


4.49 


1.5019 


4.94 


1.5974 


4.05 


1.3987 


4.50 


1.5041 


4.95 


1.5994 


4.06 


1.4012 


4.51 


1.5063 


4.96 


1.6014 


4.07 


1.4036 


4.52 


1.5085 


4.97 


1.6034 


4.08 


1.4061 


4.53 


1.5107 


4.98 


1.6054 


4.09 


1.4085 


4.54 


1.5129 


4.99 


1.6074 


4.10 


1.4110 


4.55 


1.5151 


5.00 


1.6094 


4.11 


1.4134 


4.56 


1.5173 


5.01 


1.6114 


4.12 


1.4159 


4.57 


1.5195 


5 02 


1.6134 



STEAM ENGINES 263 

Hyperbolic Logarithms of Numbers — Continued 



No. 


Log. 


No. 


Log. 


No. 


Log. 


5.03 


1.6154 


5.49 


1.7029 


5.95 


1.7834 


5.04 


1.6174 


5.50 


1.7047 


5.96 


1.7851 


5.05 


1.6194 


5.51 


1.7066 


5.97 


1.7867 


5.06 


1.6214 


5.52 


1.7084 


5.98 


1.7884 


5.07 


1.6233 


5.53 


1.7102 


5.99 


1.7901 


5.08 


1.6253 


5.54 


1.7120 


6.00 


1.7918 


5.09 


1.6273 


5.55 


1.7138 


6.01 


1.7934 


5.10 


1.6292 


5.56 


1.7156 


6.02 


1.7951 


5.11 


1.6312 


5.57 


1.7174 


6.03 


1.7967 


5.12 


1.6332 


5.58 


1.7192 


6.04 


1.7984 


5.13 


1.6351 


5.59 


1.7210 


6.05 


1.8001 


5.14 


1.6371 


5.60 


1.7228 


6.06 


1.8017 


5.15 


1.6390 


5.61 


1.7246 


6.07 


1.8034 


5.16 


1 . 6409 


5.62 


1.7263 


6.08 


1.8050 


5.17 


1.6429 


5.63 


1.7281 


6.09 


1.8066 


5.18 


1.6448 


5.64 


1.7299 


6.10 


1.83 


5.19 


1.6467 


5.65 


1.7317 


6.11 


1.8099 


5.20 


1.6487 


5.66 


1.7334 


6.12 


1.8116 


5.21 


1.6506 


5.67 


1.7352 


6.13 


1.8132 


5.22 


1.6525 


5.68 


1.7370 


6.14 


1.8148 


5.23 


1.6544 


5.69 


1.7387 


6.15 


1.8165 


5.24 


1 . 6563 


5.70 


1.7405 


6.16 


1.8181 


5.25 


1.6582 


5.71 


1.7422 


6.17 


1.8197 


5.26 


1.6601 


5.72 


1.7440 


6.18 


1.8213 


5.27 


1 . 6620 


5.73 


1.7457 


6.19 


1.8229 


5.28 


1.6639 


5.74 


1.7475 


6.20 


1.8245 


5.29 


1.6658 


5.75 


1.7492 


6.21 


1.8262 


5.30 


1.6677 


5.76 


1.7509 


6.22 


1.8278 


5.31 


1.6696 


5.77 


1.7527 


6.23 


1.8294 


5.32 


1.6715 


5.78 


1.7544 


6.24 


1.8310 


5.33 


1.6734 


5.79 


1.7561 


6.25 


1.8326 


5.34 


1.6752 


5.80 


1.7579 


6.26 


1.8342 


5.35 


1.6771 


5.81 


1.7596 


6.27 


1.8358 


5.36 


1.6790 


5.82 


1.7613 


6.28 


1.8374 


5.37 


1.6808 


5.83 


1.7630 


6.29 


1.8390 


5.38 


1.6827 


5.84 


1.7647 


6.30 


1.8405 


5.39 


1.6845 


5.85 


1.7664 


6.31 


1.8421 


5.40 


1.6864 


5.86 


1.7681 


6.32 


1.8437 


5.41 


1.6882 


5.87 


1.7699 


6.33 


1.8453 


5.42 


1.6901 


5.88 


1.7716 


6.34 


1.8469 


5.43 


1.6919 


5.89 


1.7733 


6.35 


1.8485 


5.44 


1.6938 


5.90 


1.7750 


6.36 


1.8500 


5.45 


1.6956 


5.91 


1.7760 


6.37 


1.8516 


5.46 


1.6974 


5.92 


1.7783 


6.38 


1.8532 


5.47 


1.6993 


5.93 


1.7800 


6.39 


l.«547 


5.48 


1.7011 


5.94 


1.7817 


6.40 


1.8563 



264 STEAM ENGINEERING 

Hyperbolic Logarithms of Numbers — Continued. 



No. 


Log. 


No. 


Log. 


No. 


Log 


6.41 


1.8579 


6.87 


1.9272 


7.33 


1.9920 


6.42 


1.8594 


6.88 


1.9286 


7.34 


1.9933 


6.43 


1.8610 


6.89 


1.9301 


7.35 


1.9947 


6.44 


1.8625 


6.90 


1.9315 


7.36 


1.9961 


6.45 


1.8641 


6.91 


1.9330 


7.37 


1.9974 


6.46 


1.8656 


6.92 


1.9344 


7.38 


1.9988 


6.47 


1.8672 


6.93 


1.9359 


7.39 


2.0001 


6.48 


1.8687 


6.94 


1.9373 


7.40 


2.0015 


6.49 


1.8703 


6.95 


1.9387 


7.41 


2.0028 


6.50 


1.8718 


6.96 


1.9402 


7.42 


2.0042 


6.51 


1.8733 


6.97 


1.9416 


7.43 


2.0055 


6.52 


1.8749 


6.98 


1.9430 


7.44 


2.0069 


6.53 


1.8764 


6.99 


1.9445 


7.45 


2.0082/ 


6.54 


1.8779 


7.00 


1.9459 


7.46 


2.0096 


6.55 


1.8795 


7.01 


1.9473 


7.47 


2.0109 


6.56 


1.8810 


7.02 


1.9488 


7.48 


2.0122 


6.57 


1.8825 


7.03 


1.9502 


7.49 


2.0136 


6.58 


1.8840 


7.04 


1.9516 


7.50 


2.0149 


6.59 


1.8856 


7.05 


1.9530 


7.51 


2.0162 


6.60 


1.8871 


7.06 


1.9544 


7.52 


2.0176 


6.61 


1.8886 


7.07 


1.9559 


7.53 


2.0189 


6.62 


1.8901 


7.08 


1.9573 


7.54 


2.0202 


6.63 


1.8916 


7.09 


1.9587 


7.55 


2.0215 


6.64 


1.8931 


7.10 


1.9601 


7.56 


2.0229 


6.65 


1.8946 


7.11 


1.9615 


7.57 


2.0242/ 


6.66 


1.8961 


7.12 


1.9629 


7.58 


2.0255' 
2.0268 


6.67 


1.8976 


7.13 


1.9643 


7.59 


6.68 


1.8991 


7.14 


1.9657 


7.60 


2.0281 


6.69 


1.9006 


7.15 


1.9671. 


7.61 


2.0295 


6.70 


1.9021 


7.16 


1.9685 


7.62 


2.0308 


6.71 


1.9036 


7.17 


1.9699 


7.63 


2.0321 


6.72 


1.9051 


7.18 


1.9713 


7.64 


2.0334 


6.73 


1.9066 


7.19 


1.9727 


7.65 


2.0347 


6.74 


1.9081 


7.20 


1.9741 


7.66 


2.0360 


6.75 


1.9095 


7.21 


1.9755 


7.67 


2.0373 


6.76 


1.9110 


7.22 


1.9769 


7.68 


2.0386/ 


6.77 


1.9125 


7.23 


1.9782 


7.69 


2.0399' 


6.78 


1.9140 


7.24 


1.9796 


7.70 


2.0412 


6.79 


1.9155 


7.25 


1.9810 


7.71 


2.0425 


6.80 


1.9169 


7.26 


1.9824 


7.72 


2.0438 


6.81 


1.9184 


7.27 


1.9838 


7.73 


2.0451 


6.82 


1.9199 


7.28 


1.9851 


7.74 


2.0464 


6.83 


1.9213 


7.29 


1.9865 


7.75 


2.0477 


6.84 


1.9228 


7.30 


1,9879 


7.76 


2.0490 


6.85 


1.9242 


7.31 


1.9892 


7.77 


2.0503 


6.86 


1.9257 


7.32 


1.9906 


7.78 


2.0516 



STEAM ENGINES 265 

Hyperljolic Loganthms of Numbers — Continued 



No. 


Log. 


No. 


Log. 


No. 


Log. 


7.79 


2.0528 


8.25 


2.1102 


8.71 


2.1645 


7.80 


2.0541 


8.26 


2.1114 


8.72 


2.1656 


7.81 


2.0554 


8.27 


2.1126 


8.73 


2.1668 


7.82 


2.0567 


8.28 


2.1138 


8.74 


2.1679 


7.83 


2.0580 


8.29 


2.1150 


8.75 


2.1691 


7.84 


2.0592 


8.30 


2.1163 


8.76 


2.1702 


7.85 


2.0605 


8.31 


2.1175 


8.77 


2.1713 


7.86 


2.0618 


8.32 


2.1187 


8.78 


2.1725 


7.87 


2.0631 


8.33 


2.1199 


8.79 


2.1736 


7.88 


2.0643 


8.34 


2.1211 


8.80 


2.1748 


7.89 


2.0656 


8.35 


2.1223 


8.81 


2.1759 


7.90 


2.0669 


8.36 


2.1235 


8.82 


2.1770 


7.91 


2.0681 


8.37 


2.1247 


8.83 


2.1782 


7.92 


2.0694 


8.38 


2.1258 


8.84 


2.1793 


7.93 


2.0707 


8.39 


2.1270 


8.85 


2.1804 


7.94 


2.0719 


8.40 


2.1282 


8.86 


2.1815 


7.95 


2.0732 


8.41 


2.1294 


8.87 


2.1827 


7.96 


2.0744 


8.42 


2.1306 


8.88 


2.1838, 


7.97 


2.0757 


8.43 


2.1318 


8.89 


2.1849 


7.98 


2.0769 


8.44 


2.1330 


8.90 


2.1861 


7.99 


2.0782 


8.45 


2.1342 


8.91 


2.1872 


8.00 


2.0794 


8.46 


2.1353 


8.92 


2.1883 


8.01 


2.0807 


8.47 


2.1365 


8.93 


2.1894 


8.02 


2.0819 


8.48 


2.1377 


8.94 


2.1905 


8.03 


2.0832 


8.49 


2.1389 


8.95 


2.1917 


8.04 


2.0844 


8.50 


2.1401 


8.96 


2.1928 


8.05 


2.0857 


8.51 


2.1412 


8.97 


2.1939 


8.06 


2.0869 


8.52 


2.1424 


8.98 


2.1950 


8.07 


2.0882 


8.53 


2.1436 


8.99 


2.1961 


8.08 


2.0894 


8.54 


2.1448 


9.00 


2.1972 


8.09 


2.0906 


8.65 


2.1459 


9.01 


2.1983 


8.10 


2.0919 


8.56 


2.1471 


9.02 


2.1994 


8.11 


2.0931 


8.57 


2,1483 


9.03 


2.2006 


8.12 


2.0943 


8.58 


2.1494 


9.04 


2.2017 


8.13 


2.0956 


8.59 


2.1506 


9.05 


2.2028 


8.14 


2.0968 


8.60 


2.1518 


9.06 


2.2039 


8.15 


2.0980 


8.61 


2.1529 


9.07 


2.2050 


8.16 


2.0992 


8.62 


2.1541 


9.08 


2.2061 


8.17 


2.1005 


8.63 


2.1552 


9.09 


2.2072 


8.18 


2.1017 


8.64 


2.1564 


9.10 


2.2083 


8.19 


2.1029 


8.65 


2.1576 


9.11 


2.2094 


8.20 


2.1041 


8.66 


2.1587 


9.12 


2.2105 


8.21 


"2.1054 


8.67 


2.1599 


9.13 


2.2116 


8.22 


2.1066 


8.68 


2.1610 


9.14 


2.2127 


8.23 


2.1078 


8.69 


2.1622 


9.15 


2.2138 


8.24 


2.1090 


8.70 


2.1633 


9.16 


2.2148 



^ 



266 


STEAM ENGINEERING 




Hyperbolic Logarithms of Numbers — Continued^ 

i 


No. 


Log. 


No. 


Log. 


No. 


Log. ; 


9.17 


2.2159 


9.55 


2.2565 


9.93 


2.2956 


9.18 


2.2170 


9.56 


2.2576 


9.94 


2.2966 


9.19 


2.2181 


9.57 


2.2586 


9.95 


2.2976 


9.20 


2.2192 


9.58 


2.2597 


9.96 


2.29SJ 


9.21 


2.2203 


9.59 


2.2607 


9.97 


2.2996 


9.22 


2.2214 


9.60 


2.2618 


9.98 


2.3006 


9.23 


2.2225 


9.61 


2.2628 


9.99 


2.3016: 


9.24 


2.2235 


9.62 


2.2638 


10.00 2.3026 


9.25 


2.2246 


9.63 


2.2649 


10.25 


2.3279: 


9.26 


2.2257 


9.64 


2.2659 


10.50 


2.3513 i 


9.27 


2.2268 


9.65 


2.2670 


10.75 


2.3749 1 


9.28 


2.2279 


9.66 


2.2680 


11.00 


2.3979 


9.29 


2.2289 


9.67 


2.2690 


11.25 


2.4201 : 


9.30 


2.2300 


9.68 


2.2701 


11.50 


2.4430 


9.31 


2.2311 


9.69 


2.2711 


11.75 


2.4636 


9.32 


2.2322 


9.70 


2.2721 


12.00 


2.4849 


9.33 


2.2332 


9.71 


2.2732 


12.25 


2.5052 1 


9.34 


2.2343 


9.72 


2.2742 


12.50 


2.5262 


9.35 


2.2354 


9.73 


2.2752 


12.75 


2.5455 


9.36 


2.2364 


9.74 


2.2762 


13.00 


2.5649 


9.37 


2.2375 


9.75 


2.2773 


13.25 


2.5840 


9.38 


2.2386 


9.76 


2.2783 


13.50 


2.6027 


9.39 


2.2396 


9.77 


2.2793 


13.75 


2.6211 


9.40 


2.2407 


9.78 


2.2803 


14.00 


2.6391 


9.41 


2.2418 


9.79 


2.2814 


14.25 


2.6567 


9.42 


2.2428 


9.80 


2.2824 


14.50 


2.6740 


9.43 


2.2439 


9.81 


2.2834 


14.75 


2.6913 


9.44 


2.2450 


9.82 


2.2844 


15.00 


2.7081 


9.45 


2.2460 


9.83 


2.2854 


15.50 


2.7408 


9.46 


2.2471 


9.84 


2.2865 


16.00 


2.7726 


9.47 


2.2481 


9.85 


2.2875 


16.50 


2.8034 


9.48 


2.2492 


9. 86 


2.2885 


17.00 


2.8332 


9.49 


2.2502 


9.87 


2.2895 


17.50 


2.8621 


9.50 


2.2513 


9.88 


2.2905 


18.00 


2.8904 


9.51 


2.2523 


9.89 


2.2915 


18.50 


2.9173 


9.52 


2.2534 


9.90 


2.2925 


19.00 


2.9444 


9.53 


2.2544 


9.91 


2.2935 


19.50 


2.9703 


9.54 


2.2555 


9.92 


2.2946 


20.00 


2.9957 



In order to save the trouble of applying 
a rule or a formula for the purpose of 
determining the mean effective pressure 
for stated conditions, the next table is 



STEAM ENGINES 267 

given, but the following explanation 
must be carefully noted, as otherwise 
correct results will not be secured from 
it. 

The first horizontal line contains the 
various points at which steam is cut off 
from the cylinder, ranging from 1/10 to 
9/10 of the stroke. The second line is the 
corresponding ratio of expansion, pro- 
vided that clearance is not taken into 
accoimt. The first column contains 
initial pressures from 10 to 200 pounds, 
all absolute or above a perfect vacuum, 
therefore, when taking gauge pressure 
for an example in practice at or near the 
sea level, 15 pounds should be added to 
it, and the resulting number found in 
the first column. Gauge pressure might 
have been used in place of absolute 
pressures, but in that case it would only 
apply to places where the atmosphere 
weighs nearly 15 pounds to the square 
inch, whereas the plan adopted makes it 
available for all conditions. For the 
same reason the back pressure was not 
subtracted, as there is no universal rule 
to follow. 

To fully illustrate the use of this table 
of "Average Pressure of Steam" suppose 
that the initial pressure as shown by 



268 STEAM ENGINEERING 

measuring upward from the atmos- ;[ 
pheric line of an indicator diagram, is 
105 pounds, the atmospheric pressure is 
practically 15 pounds, and the point of 
cut off is at M stroke. The engine is 
run non-condensing. What is the mean \ 
effective pressure? 

The absolute pressure is 105+15 = 120 
poimds. Finding 120 in the first col- 
umn and following the line to the right 
hand to the sixth column under 3^ we 
find that the average pressure is 71.58 
pounds. Subtracting 15 for the atmos- 
phere and one for friction of steam, or 
16 pounds back pressure, shows that 
under these conditions the mean effec- 
tive pressure is 55.58 pounds. This 
agrees with the result secured by calcu- 
lation as demonstrated on preceding 
pages. 

Assuming that a condensing engine is 
run under conditions that agree with the 
preceding example, except that while 
the load is increased enough to maintain 
the cut off at 3^ stroke, the back pres- 
sure is reduced to 3 pounds. What is 
the mean effective pressure? 

We have demonstrated that the aver- 
age pressure is 71.58 pounds, therefore 



STEAM ENGINES 269 

the mean effective pressure is 71.58 — 3 
= 68.58 pounds. 

Assuming that the non-condensing 
engine has 7 per cent, of clearance, 
what is the mean effective pressure, pro- 
vided the clearance is taken into consid- 
eration? 

Referring to the table entitled "Actual 
Ratios of Expansion," and following 
down the first column under P of C until 
.25 is found (meaning that the point of 
cut off is at ]4: or .25 of the stroke), then 
taking the horizontal line until the col- 
umn under 7 per cent, is reached, we find 
that the actual ratio of expansion is 3.34. 
In the table of "Average Pressure of 
Steam" the nearest ratio of expansion is 
3.33 which practically agrees with the 
conditions above mentioned. In this 
column on the horizontal line beginning 
with 120 poimds pressure we learn that 
the average pressure is 79.31 pounds 
from which must be subtracted the 
back pressure of 16 pounds, therefore 
the mean effective pressure is 63.31 
pounds. 

Suppose that a condenser is added to 
this engine, and more machinery is in- 
stalled in the mill or shop, causing the 
cut off to take place at 1/5 or .20 of the 



270 STEAM ENGINEERING 

Stroke, the clearance remaining at 7 per 
cent., what is the mean effective pres- 
sure? 

Following the directions above given 
the actual ratio of expansion is 3.96, 
which is 4 for all practical purposes. 
With 120 pounds initial pressure, and 4 
expansions, or a ratio of expansion of 4, 
the average pressure is 71.58 pounds and 
the mean effective pressure is 71.58—3 
= 68.58 pounds. 

The following explanation of the next 
table is given in condensed form for the 
convenience of application: 

Cut off = The proportion of the 
stroke completed by the 
piston when the cut off 
valve closes. 

R of E = Ratio of expansion, or 1 
divided by the "cut off" 
as above described. 

I P = Initial pressure absolute, 
or the pressure at the 
beginning of the stroke 
above a vacuum, which 
is maintained until the 
cut off valve closes. 



STEAM ENGINES 271 

AVERAGE PRESSURE OF STEAM 



Cut-off 


Vio 


H 


Va 


\^ 


M 


R.ofE. 


10 


8 


6 


5 


4 


I. P. 


Average Pressure 



3.30 
4.94 
6.60 
8.25 
9.90 
11.56 
13.21 
14.86 
16.51 
18.16 
19.81 
21.46 
23.10 
24.77 
26.42 
28.07 
29.73 
31.37 
33.03 
34.68 
36.38 
37.98 
39.64 
41.28 
42.92 
44.59 
46.23 
47.89 
49.54 
51.19 
52.84 
54.49 
56.15 
57.80 
59.45 
61.10 
62.75 
64.40 
66.06 



3.84 
5 77 
1.-' 
9.62 
11.54 
13.47 
15.39 
17.32 
19.24 
21.16 
23.09 
25.01 
26.94 
28.86 
30.79 
32.71 
34.64 
35.56 
38.48 
40.41 
42.34 
44.26 
46.18 
48.10 
50.03 
51.95 
53.88 
55.80 
57.73 
59.65 
61.58 
63.50 
65.42 
67.35 
69.27 
71.20 
73.12 
75.05 
76.97 



4.64 
6.98 
9.29 
11.63 
13.96 
16.28 
18.59 
20.94 
23.26 
25.59 
27.92 
30.24 
32.55 
34.90 
37.21 
39.55 
41.88 
44.20 
46.51 
48.56 
51.17 
53.51 
55.84 
58.16 
60.47 
62.82 
65.13 
67.47 
69.80 
72.13 
74.45 
76.78 
79.10 
81.43 
83.76 
86.08 
88.41 
90.74 
93.05 



5 


21 


7 


82 


10 


43 


13 


04 


15 


65 


18 


26 


20 


87 


23 


48 


26 


09 


28 


69 


31 


31 


33 


91 


36 


53 


39 


13 


41 


74 


44.351 


46 


96| 


49 


57 


52 


18 


54 


78 


57 


40 


60 


00 


62 


62 


65 


22 


67 


83 


70 


40 


73 


05 


75 


66 


78 


27 


80 


87 


83 


48 


86 


09 


88.701 


91 


31 


93 


771 


96 


53 


99 


14 


101 


7 


104 


4 



5.96 

8.94 

11.93 

14.91 

17.90 

20.87 

23.86 

26.84 

29.82 

32.80 

35.79 

38.77 

41.75 

44.73 

47.72 

50.57 

53.68 

56.66 

59.65 

62.63 

65.61 

68.59 

71.58 

74.56 

77.54 

80.52 

83.51 

86.49 

89.47 

92.45 

95.44 

98.42 

101.4 

103.3 

107.3 

110.3 

113.3 

116.3 

119.3 



r 1 

272 STEAM ENGINEERING 

AVERAGE PRESSURE OF STEAM 



Cut-off 


%o 


Vs 


%o 


^ 


%o 


% 


R.ofE. 


3.33 


3 


2.5 


2 


1.66 


1.5 


LP. 


Average Pressure. 


10 


6.60 


6.98 


7.66 


8.46 


9.03 


9.36 


15 


9.91 


10.49 


11.49 


12.69 


13.55 


14.05 


20 


13.21 


13.98 


15.32 


16.93 


18.07 


18.73 


25 


16,52 


17.48 


19.16 


21.16 


22.59 


23.41 


30 


19.82 


20.99 


22.99 


25.39 


27.11 


28.10. 


35- 


23.13 


24.48 


26.82 


29.64 


31.62 


32.78 


40 


26.43 


27.98 


30.65 


33.86 


36.15 


37.43 


45 


29.74 


31.48 


34.38 


38.09 


40.66 


42.14 


50 


33.04 


34.96 


38.32 


42.32 


45.18 


46.82 


55 


36.38 


38.48 


42.15 


46.55 


49.93 


51.51 


60 


39.65 


41.98 


45.98 


50.79 


54.22 


56.20 


65 


44.00 


45.47 


49.81 


55.02 


59.00 


60.88 


70 


46.26 


48.96 


53.65 


59.25 


63.26 


65.56 


75 


49.61 


52.47 


57.48 


63.48 


68.08 


70.25 


80 


52.87 


55.96 


61.31 


67.72 


72.30 


74.86 


85 


56.23 


59.47 


65.14 


71.95 


77.16 


79.61 


90 


59.48 


62.97 


68.97 


76.18 


81.34 


84.28 


.95 


62.84 


66.46 


72.80 


80.41 


86.24 


88.98 


100 


66.08 


69.92 


76.64 


84.65 


90.36 


93.64 


105 


69.46 


73.46 


80.47 


88.88 


95.32 


98.35 


110 


72.70 


76.94 


84.30 


93.11 


99.40 


103.0 


115 


76.07 


80.46 


88.13 


97.34 


104.4 


107.7 


120 


79.31 


83.96 


91.96 
95.80 


101.6 


108.4 


112.4 


125 


82.69 


87.45 


105.8 


113.4 


117.0 


130 


85.91 


90.92 


99.63 


110.0 


117.5 


121.7 


135 


89.31 


94.45 


103.4 


114.2 


122.5 


126.0 


140 


92.52 


97.94 


107.3 


118.5 


126.5 


131.1 


145 


95.92 


101.4 


111.1 


122.7 


131.6 


135.8 


150 


99.13 


104.9 


114.9 


126.9 


135.5 


140.5 


155 


102.5 


108.4 


118.7 


131.2 


140.7 


145.5 


160 


105.8 


111.9 


122.6 


135.4 


145.2 


149.9 


165 


109.1 


115.4 


126.4 


139.6 


149.7 


154.5 


170 


112.4 


118.9 


130.2 


143.9 


154.3 


159.2 


175 


115.7 


122.6 


134.1 


148.1 


158.9 


163.9 


180 


119.0 


125.9 


137.9 


152.3 


163.4 


168.6 


185 


122.3 


129.4 


141.7 


156.6 


167.9 


173.2 


190 


125.6 


132.9 


145.6 


160.8 


172.4 


177.3 


195 


129.0 


136.4 


149.4 


165.0 


177.0 


182.6 


200 


132.3 


139.9 


153.3 


169.3 


180.7 


187.3 



STEAM ENGINES 273 

AVERAGE PRESSURE OF STEAM 



Cut-off 


%o 


H 


«/io 


H 


%o 


R.ofE. 


1.43 


1.33 


1.25 


1.14 


1.11 


LP. 


Average Pressure 


10 


9.48 


9.65 


9.78 


9.92 


9.94 


15 


12.23 


14.48 


14.67 


14.88 


14.92 


20 


18,98 


19.30 


19.56 


19.84 


19.89 


25 


23.72 


24.13 


24.46 


24.80 


24.86 


30 


28.46 


28.96 


29.35 


29.76 


29.84 


35 


33.21 


33.79 


34.24 


34.72 


34.81 


40 


37.95 


38.61 


39.13 


39,68 


39.78 


45 


42.70 


43.44 


44.02 


44,64 


44.75 


50 


47.44 


48.27 


48.92 


49.60 


49.73 


55 


52.19 


53.09 


53.81 


54.56 


54.70 


60 


56.93 


57.92 


58.70 


59.52 


59.67 


65 


61.68 


62.75 


63.59 


64.48 


64.64 


70 


66.42 


67.57 


68.48 


69.44 


69.62 


75 


71.17 


72.40 


73.38 


74.40 


74.59 


80 


75.91 


77.23 


78.27 


79.37 


79.56 


85 


80.66 


82.06 


83.16 


84,32 


84.54 


90 


85.40 


86.88 


88.05 


89.28 


89.50 


95 


90.15 


91.71 


92.94 


94.25 


94.48 


100 


94.89 


96.54 


97.84 


99.21 


99.46 


105 


99.63 


101.3 


102.7 


104.1 


104.4 


110 


104.4 


106.2 


107.6 


109.1 


109.4 


115 


109.1 


111.0 


112.5 


114.0 


114.3 


120 


113.8 


115.8 


117.4 


119.0 


119.3 


125 


118.6 


120.6 


122.3 


124.0 


124.3 


130 


123.3 


125.5 


127.2 


128.9 


129.3 


135 


128.1 


130.3 


132.0 


133.9 


134.2 


140 


132.8 


135.1 


136.9 


138.8 


139.2 


145 


137.5 


139.9 


141.8 


143.8 


144.2 


150 


142.3 


144.8 


146.7 


148.8 


149.2 


155 


147.0 


149.6 


151.6 


153.7 


154.1 


160 


151.8 


154,4 


156.5 


158.7 


159.1 


165 


156.5 


159.2 


161.4 


163.6 


164.1 


170 


161.3 


164.1 


166.3 


168.6 


169.0 


175 


166.0 


168.9 


171.2 


173.6 


174.0 


180, 


170.8 


173.7 


176.1 


178.5 


179.0 


185 


175.5 


178.6 


181.0 


183,5 


184.0 


190 


180.3 


183,4 


185.9 


188.5 


188.9 


195 


185.0 


188.2 


192.3 


193.4 


193.9 


200 


189.8 


193.0 


195.6 


198.4 


198.9 



274 STEAM ENGINEERING 

CAUTION 

Do not attempt to use the foregoing 
table until you have carefully read the 
explanation of it which precedes the 
table, as otherwise correct results may 
not be obtained owing to misapplication 
of the figures given. 

As a further illustration of its value 
take the case of an engine located on 
very high ground where the atmosphere 
pressure is only about two-thirds of 
what we find it at sea level, or say 10 
pounds, for convenience. If the steam 
gauge indicates 110 pounds and the initial 
pressure is 5 pounds less, the initial pres* 
sure absolute is 110—5 + 10 = 115 
pounds. Assuming that the steam is 
cut off at 3^ stroke and ignoring the 
clearance, the table indicates that the 
average pressure is 68.59 pounds. If this 
engine is operated without a condensei 
the mean effective pressure is 68.5ti 
— 16 = 52.59 pounds. If the condensei 
is used it is 68.59 — 3=65.59 pounds. 

REMARKS ON CALCULATING THE HORSE 
POWER OF STEAM ENGINES 

Preceding calculations on this impor- 
tant subject relate to simple double 
acting engines, which constitute a ma- 
jority of the engines now in use, but 



STEAM ENGINES 275 

there are others in service, and as the 
proper way to determine the power they 
develop may not be plain to the young 
engineer, explanations are herewith ' 
given that will prove useful in such cases. 

Where a single acting engine is in use 
the effective piston speed is only one- 
half of what it is for a double acting 
engine, therefore the indicated horse 
power will be but one-half of that ob- 
tained by preceding rules and formulas. 

Where there are two single acting 
cylinders the engine is to be treated as 
if it were a double acting engine with one 
cylinder. When calculating the power 
developed by these engines, the piston 
speed is taken while steam is acting on 
the piston. If this principle is always 
applied, correct results will be obtained. 
When treating a double acting engine, 
the total piston speed is to be used, but 
the mean effective pressure for two dia- 
grams should not be added together and 
the result used in making the calculation, 
as that gives twice as much power as 
really is developed. It is proper to add 
them but the sum must be divided 
by 2. 

A double engine is to be treated as if it 
consisted of two separate engines located 
in different parts of the works .where 
machinery is used. The power of each 
is to be computed, in accordance with 



276 STEAM ENGINEERING 

rules given in this book, and the results 
added together. 

Where great accuracy is required it 
becomes necessary to make allowance 
for the space occupied by the piston rod 
on the area of the piston, because steam 
pressure cannot act on this space. In 
the case of a single acting engine with 
two cylinders, no allowance is to be 
made for the rod as steam only acts on 
the head end, where there is no rod to 
reduce the area. 

With a double acting engine (not 
fitted with a tail rod), it is proper to sub- 
tract one-half of the area of the piston 
rod, and call the remainder the effective 
area of the piston. The piston rod occupies 
space on one side of the piston, while the 
other side is clear, therefore if one-half 
of it is assumed to be on each side of 
it, the result is correct. 

For example, take a 20= inch piston 
with a piston rod S'J/g inches in diameter. 
The area of a 20= inch circle is 314.16 
square inches, and the area of a 3J^ 
circle is 11.793 square inches. Then 
(314.16 - 11.793) -f 314.16 -v- 2 = 308.26 
square inches. The area of the piston 
rod is 11.793 square inches, one-half of 
which is 5.896; then 314.16-5.896 = 
308.26 square inches as before. 

Large horizontal engines are fitted 
with heavy pistons that rest on the bot- 



STEAM ENGINES 277 

torn of the cylinders and cause much 
friction. Unless superior lubrication is 
provided, the wear is excessive, and in 
order to improve these conditions, a tail 
rod is added which is illustrated in Fig. 
41. The piston rod is continued through 
the cylinder head on the head end, 
which is fitted with a stuffing box sim- 
ilar to the crank end cylinder head. The 
outer end of this piston rod is carried by 
a second crosshead that travels on a 
lower guide, and as this is always in 
sight, it can be properly lubricated, thus 
preventing imdue wear on the piston and 
the cylinder. 

When calculating the power of such 
engines the whole area of the rod should 
be subtracted from the area of the piston. 
For a 20-inch piston with a rod 3J^ 
inches in diameter the effective area is 
314.16-11.793=302.36 square inches. 

Where several pounds back pressure 
are added to an engine in order to use 
the exhaust steam for heating a mill or 
shop, or for any other purpose, it in- 
creases the load and makes it necessary 
to bum more coal. As this change 
makes no difference in the indicated 
horse power, it does not seem to be con- 
sistent to some engineers, but it admits 
of a perfectly logical explanation. The 
indicated horse power of an engine is 
represented by the area of the indicator 



278 



STEAM ENGINEERING 




STEAM ENGINES 279 

diagram taken from it, and while this 
varies with changes made in the ma- 
chinery driven by this engine, it is not 
affected by changes in the back pressure. 
How then is the apparent change in the 
load accounted for? 

Careful study of the following state- 
ments will make this clear. Taking for 
standard conditions an engine in which 
there is no back pressure above the atmos- 
phere, the steam and expansion lines 
average a certain number of inches 
above the vacuum line, and this height 
represents the total load on the engine. 
If back pressure is added to the piston 
by forcing exhaust steam around the 
shop until the back pressure line is raised 
a certain part of an inch, the average 
height of the steam and expansion lines 
will be raised the same fraction of an 
inch, showing that the total load has 
been increased. 

On the other hand, if a condenser is 
added and the back pressure line is 
lowered a certain fraction of an inch by 
it, the average height of the steam and 
the expansion lines will be lowered the 
same fraction of an inch, showing that 
the total load has been decreased, and 
the coal required to generate steam to 
operate the same machinery is less. 
Under all of the foregoing changes it is 
assumed that the same amount of power 



280 STEAM ENGINEERING 

is required to drive the works, as any ; 
change in this respect will prevent 
proper comparison of the effects of the 
above mentioned difference in back 
pressure. 

HORSE POWER CONSTANTS 

It is frequently necessary to deter- 
mine the power that an engine is devel- 
oping under varying conditions, which 
cause changes in the mean effective 
pressure. As the speed is assumed to be 
constant under these changes (and it is 
practically so with an up-to-date engine), 
the mean effective pressure is the only 
factor that changes, hence it follows 
that if a constant is determined which 
represents the power developed for one 
pound mean effective pressure, it is only 
necessary to multiply this constant by 
the mean effective pressure to determine 
the power developed. This constant is 
found by use of the following rule. 

Multiply the area of the piston by the 
piston speed in feet per minute, and 
divide by 33,000. The quotient is the 
horse power constant for that engine 
under given conditions. 

For illustration suppose that an en- 
gine is fitted with a piston 20 inches in 
diameter, the area of which is 314.16 
square inches. The stroke is 42 inches 
and the speed 90 revolutions per minute, 



STEAM ENGINES 281 

giving a piston speed of 630 feet. Then 
314.16X630^-33,000 = 5.9976 which is 
the horse power constant. If the mean 
effective pressure is 50 pounds to the 
square inch, it develops 5.9976X50 = 
299.88 horse power. 

The next table gives the horse power 
constants for cylinders from 4 to 60 
inches in diameter, with piston speed 
from 10 to 900 feet per minute. To illus- 
trate its practical application, suppose 
that a 24X48 inch engine, at 100 revo- 
lutions per minute shows a mean effec- 
tive pressure of 56 pounds; what power 
is developed? 

The piston speed is 8 feet per revolu- 
tion, or 800 feet per minute. Finding 
the column under 800 and following it 
downward until the horizontal line 
opposite 24 is found, the horse power 
constant is 10.967. Then 10.967X56 = 
614.15 horse power developed under 
these conditions. 

Suppose that an 18X36 inch engine 
revolves 80 times per minute, and indi- 
cator diagrams from it show 47 pounds 
mean effective pressure; what power 
is developed? 

The piston speed is 6 feet per revolu- 
tion, or 480 feet per minute. In the 
column under 400 and on the horizontal 
line opposite 18 the horse power constant 



282 



STEAM ENGINEERING 



is 3.0845. On the same line under 80 I 
the constant is .6169. These two con- ' 
St ants are to be added together to ac- 
count for 480 feet per minute, and 
3.0845 + .6169 = 3.7014. Multiplying 
this by 47 shows that 173.96 horse 
power was developed when the diagrams 
were taken. 



HORSE POWER CONSTANTS 







Piston 


speed per minute. 




s 

.2 

o 


10 


20 


30 


40 


50 


60 


4 


.0038 


.0076 
.0119 


.0114 


.0152 


.0190 


.0228 


5 


.0059 


.0178 


.0238 


.0297 


.0357 


6 


.0085 


.0171 


.0257 


.0342 


.0428 


.0514 


7 


.0116 


.0233 


.0349 


.0466 


.0583 


.0699 


8 


.0152 


.0304 


.0457 


.0609 


.0761 


.0913 


9 


.0192 


.0385 


.0578 


.0771 


.0963 


.1156 


10 


.0238 


.0476 


.0714 


.0952 


.1190 


.1428 


11 


.0288 


.0576 


.0863 


.1151 


.1439 


.1729 


12 


.0342 


.0685 


.1028 


.1370 


.1713 


.2056 


13 


.0402 


.0804 


.1206 


.1608 


.2011 


.2413 


14 


.0466 


.0933 


.1399 


.1865 


.2332 


.2798 


15 


.0535 


.1071 


.1606 


.2142 


.2677 


.3213 


16 


.0609 


.1218 


.1827 


-.2437 


.3046 


.3655 


17 


.0687 


.1275 


.1963 


.2651 


.3339 


.4026 


18 


.0771 


.1542 


.2313 


.3084 


.3855 


.4626 


19 


.0859 


.1718 


.2577 


.3436 


;4295 


.5155 


20 


.0952 


.1904 


.2856 


.3808 


.4760 


.5712 


21 


.1049 


.2099 


.3148 


.4198 


.5247 


.6297 


22 


.1151 


.2303 


.3455 


.4607 


.5759 


.6911 


23 


.1259 


.2518 


.3777 


.5036 


.6295 


.7554 


24 


.1370 


.2741 


.4112 


.5483 


.6854 


.8225 


25 


.1487 


.2975 


.4462 


.5950 


.7437 


.8925 


26 


.1608 


.3217 


.4826 


.6435 


.8044 


.9653 


27 


.1735 


.3470 


.5205 


.6940 


.8675 


1.0410 


28 


.1865 


.3731 


.5597 


.7463 


.9329 


1.1196 


29 


.2001 


.4003 


.6004 


.8006 


1.0008 


1.2009 


30 


.2142 


.4284 


.6426 


.8568 


1.0710 


1.2852 


31 


.2287 


.4574 


.6861 


.9148 


1.143.6 


1 .3723 


32 


.2437 


.4874 


.7311 


.9748 


1.2186 


1.4623 


33 


.2591 


.5183 


.7775 


1.0367 


1.2959 


1.5551 


34 


.2751 


.5502 


.8253 


1.1005 


1.3766 


1.6508 



STEAM ENGINES 283 

Horse Power Constants — Continued. 
Piston speed per minute. 



.1 


10 


20 


30 


40 


50 


60 


85 


.2915 


.5831 


.8746 


1.1662 


1.4578 


1.7493 


36 


,3084 


.6169 


.9253 


1.2338 


1.5422 


1.8507 


87 


.8258 


.6516 


.9774 


1.3033 


1.6291 


1.9549 


8S 


.8486 


.6873 


1.0310 


1.3747 


1.7184 


2.0620 


89 


.3620 


.7240 


1.0860 


1.4480 


1.8100 


2.1720 


40 


.8808 


.7616 


1.1424 


1.5232 


1.9040 


2.2848 


41 


.4000 


.8001 


1.2002 


1.6003 


2.0004 


2.4005 


42 


.4198 


.8386 


1.2585 


1.6783 


2.0982 


2.5180 


48 


.4400 


.8801 


1.3202 


1.7602 


2.2003 


2.6404 


44 


.4607 


.9215 


1.3823 


1.8431 


2.3038 


2.7646 


45 


.4819 


.9639 


1.4459 


1.9278 


2.4098 


2.8917 


46 


.50.86 


1.0072 


1.5108 


2.0144 


2.5180 


3.0216 


47 


.5257 


1.0515 


1.5772 


2.1030 


2.6287 


3 . 1545 


48 


.548.8 


1.0967 


1.6451 


2.1934 


2.7418 


3.2901 


49 


.5714 


1 . 1429 


1.7143 


2.2858 


2.8572 


3.4286 


50 


.5950 


1.1900 


1.7850 


2.3800 


2.9750 


3.5700 


51 


.6190 


1.2381 


1.8571 


2.4762 


3.0952 


3.7142 


52 


. 6485 


1.2871 


1.9307 


2.5742 


3.2178 


3.8613 


58 


.6685 


1.3371 


2.0036 


2.6742 


3.3427 


4.0113 


54 


.6940 


1.3880 


2.0820 


2.7760 


3.4700 


4.1640 


55 


.7199 


1.4399 


2.1599 


2.8798 


3.5998 


4.3197 


56 


.7463 


1.4927 


2.2391 


2.9855 


3.7318 


4.4782 


57 


.7782 


1.5465 


2.3198 


3.0930 


3.8663 


4.6396 


58 


.8006 


1.6013 


2.4019 


3.2025 


4.0032 


4.8038 


59 


.8284 


1.6570 


2.4854 


3.3139 


4 . 1424 


4.9709 


60 


.8568 


1.7136 


2.5704 


3.4272 


4.2840 


5.1408 



Piston speed per minute. 



i 


70 


80 


90 


100 


200 


300 


4 


.0266 


.0304 


.0342 


.0381 


.0762 


.1142 


5 


.0416 


.0476 


.0535 


.0595 


.1190 


.1785 


6 


.0599 


.0685 


.0771 


.0857 


.1714 


.2570 


7 


.0816 


.0933 


.1049 


.1166 


.2332 


.3499 


8 


.1066 


.1218 


.1370 


.1523 


.3046 


.4570 


9 


.1849 


.1542 


.1735 


.1928 


.3856 


.5783 


10 


.1666 


.1904 


.2142 


.2380 


.4760 


.7140 


11 


.2015 


.2303 


.2581 


.2880 


.5760 


.8639 


12 


.2899 


.2741 


.3084 


.3427 


.6854 


1.0282 


18 


.2815 


.3217 


.3620 


.4022 


.8044 


1.2067 


14 


.3265 


.3731 


.4198 


.4665 


.9330 


1.3994 


15- 


.3748 


.4284 


.4819 


.5355 


1.0710 


1.6065 



284 



STEAM ENGINEERING 



Horse Power Constants — Continued. 



Piston speed per minute. 



5 


70 


80 


90 


100 


200 


300 


16 


.4265 


.4874, .5483 


.6093!l.2186 


1.8278 


17 


.4614 


.5402 


.6190 


.6878 


1.2756 


1.9635 


18 


.5397 


.6169 


.6940 


.7711 


1.5422 


2.3134 


19 


.6014 


.6873 


.7732 


.8592 


1.7184 


2.5775 


20 


.6664 


.7616 


.8568 


.9520 


1.904012.8560 


21 


.7347 


.8396 


.9446 


1.0496 


2.099213.1488 


22 


.8063 


.9215 


1.0367 


1.1519 


2.3038 


3.4558 


23 


.8813 


1.0072 


1.1331 


1.2590 


2.5180 


3.7771 


24 


.9516 


1.0967 


1.2338 


1.3709 


2.7418 


4.1126 


25 


1.0413 


1.1900 


1.3388 


1.4875 


2.9750 


4.4625 


26 


1.1262 


1.2871 


1.4480 


1 . 6089 


3.217814.8266 


27 


1.2145 


1.3880 


1.5615 


1.7350 


3.4700l5.2051 


28 


1.3061 


1.4927 


1.6793 


1.8659 


3.731815.5978 


29 


1.4011 


1.6013 


1.8014 


2.0016 


4.0032 


6.0047 


30 


1.4994 


1,7136 


1.9278 


2.142014.2840 


6.4260 


31 


1.6010 


1.8297 


2.0585 


2.2872 


4.5744 


6.8615 


32 


1.7060 


1.4497 


2.1934 


2.4371 


4.8742 


7.3114 


33 


1.8143 


2.0735 


2.3326 2.5918 


5.1836 


7.7755 


34 


1.9259 


2.2010 


2.4762 


2.7513 


5.5026 


8.2538 


35 


2.0409 


2.3324 


2.6240 


2.9155 


5.8310 


8.7465 


36 


2.1591 


2.467612.7760 


3.0845 


6.1690 


9.2534 


37 


2.2808 


2.6066 


2.9324 


3.2582 


6.5164 


9.7747 


38 


2.4057 


2.7494 


3.0930 


3.4367 


6.8734 


10.310 


39 


2.5340 


2.8960 


3.2580 


3.6200 


7.2400 10.860 


40 


2.6656 


3.0464 


3.4272 


3.8080 


7.6160 11.424 


41 


2.8005 


3.2006 


3.6007 


4.0008 


8.0016 12.002 


42 


2.9378 


3.3577 


3.7775 


4.1983 


8.3866 12.585 


43 


3.0804 


3.5205 


3.9606 


4.4006 


8.8012 13.202 


44 


3.2254 


3.6861 


4.1469 


4.6077 


9.2154 13.823 


45 


3.3737 


3.8556 


4.3376 4.8195 


9.6390 14.459 


46 


3.5253 


4.0289 


4.5325 5.0361 


10.0721 15. 108 


47 


3.6802 


4.2059 


4.7317 


5.2574 


10.515 15.772 


48 


3.8385 


4.3868 


4.9352 


5.4835 


10.967 


16.451 


49 


4.0001 


4.5715 


5.1429 


5.7144 


11.429 


17.143 


50 


4.1650 


4.7600 


5.3550 


5.9500 


11.900 


17.850, 


51 


4.3333 


4.9523 


5.5713|6.1904 


12.381 


18.571" 


52 


4.5049 


5.1484 


5.7920 6.4355 


12.871 


19.307 


53 


4.6790 


5.3483 


6.0169 6.6854 


13.371 


20.056 


54 


4.8581 


5.5521 


6.2461 6.9401 


13.880 


20.820 


55 


5.0397 


5.7596 


6.4796 7.1995 


14.399 


21.599 


56 


5.2246 


5.9709 


6.71737.4637 


14.297 


22.391 


57 


5.4128 


6.1861 


6.9584 7.7326 


15.465 


23.198 


58 


5 . 6044 


6.4051 


7.205718.0063 


16.013 


24.019 


59 


5.7993 


6.6278 7. 4563!8. 2849 


16.570 


24.854 


60 


5.9976 


8. 854417. 7112|8. 5680 17.136 


25.704 



STBAM ENGINES 285 

Horse Power Constants— Ccw/tn«*(i 



Piston speed per minute. 


1 

a 


400 


500 


600 


700 


800 


900 


4 


.15231 .1904 


.2285 


.2666 


.3046 


.3427 


5 


.2380 


.2975 


.3570 


.4165 


.4760 


.5355 


6 


.3427 


.4284 


.5141 


.5998 


.6854 


.7711 


7 


.4665 


.5831 


.6997 


.8163 


.9330 


1.049S 


8 


.6093 


.7616 


.9139 


1.0662 


1.2186 


1.3709 


9 


.7711 


.9639 


1.1567 


1.3495 


1.5422 


1.7350 


10 


.9520 


1.1900 


1.4280 


1.6660 


1.9040 


2.1420 


11 


1.1519 


1.4399 


1.7279'2.0159 


2.3038 


2.5818 


12 


1.3709 


1.7136 


2.0563 


2.3990 


2.7418 


3.0845 


13 


1.6089 


2.0111 


2.4133 


2.8155 


3.2178 


3.6200 


14 


1.8659 


2.3324 


2.7989 


3.2654 


3.7318 


4.1983 


15 


2.1420 


2.6775 


3.2130 


3.7485 


4.2840 


4.8195 


IG 


2.4371 


3.0464 


3.6557 


4.2650 


4.8742 


5.4835 


17 


2.6513 


3.3391 


4.0269 


4.6147 


5.4026 


6.1904 


18 


3.0845 


3.8556 


4 . 6267 


5.3978 


6.1690 


6.9401 


19 


3.4367 


4.2959 


5.1551 


6.0143 


6.8734 


7.7326 


20 


3.8080 


4.7600 


5.7120 


6.6640 


7.6160 


8.5680 


21 


4.1983 


5.2479 


6.2975 


7.3471 


8.3966 


9.4462 


22 


4.6077 


5.7596 


6.9115 


8.0634 


9.2154 


10.367 


23 


5.0361 


6.2951 


7.5541 


8.8131 


10.072 


11.331 


24 


5.4835 


6.8544 


8.2253 


9.5962 


10.967 


12.338 


25 


5.9500 


7.4375 


8.9250 


10.413 


11.900 


13.388 


26 


6.4355 


8.0444 


9.6534 


11.262 


12.871 


14.480 


27 


6.9401 


8.6751 


10.410 


12.145,13.880 


15.615 


28 


7.4637 


9.3296 


11.196 


13. 061114. 927 


16.793 


29 


8.0063 


10.008 


12.009 


14.011 16.013 


18.014 


30 


8.5680 


10.710 


12.852 


14.994 17.136 


19.278 


31 


9.1487 


11.436 


13.723 


16.010 18.297 


20.585 


32 


9.7485 


12.186 14.623 


17.060 19.497 


21.934 


33 


10.367 


12.959 


15.551 


18.143 


20.735 


23.326 


34 


11.005 


13.756 


16.508 


19.259 


22.010 24.762 


35 


11.662 


14.578 


17.493 


20.409 


23.324126.240 


36 


12.338 


15.422 


18.507 


21.591 


24.676127.760 


37 


13.033 


16.291 


19.549 


22.808 26.066,29.324 


38 


13.747 


17.184 


20.620 


24.057 27. 494130. 930 


39 


14.480 


18.100 


21.720 


25.340 


28. 960132. 580 


40 


15.232 


19.040 


22.848 


26.656 


30.464 34.272 


41 


16.003 


20.004 


24.005 


28.005 


32.006 36.007 


42 


16.783 


20.982 


25.180 


29.378 


33.577| 


37.775 


43 


17.602 


22.003 


26.404 


30.804 


35.205 


39.606 


44 


18.431 


23.038 27.646 


32.254 


36.861 


41.469 


45 


19.278 


24,098 28.917 


33.737 


38.556 


43.376 


46 


20.144 


25.180 30.216 


35.253 


40.289145.325 

i 



286 



STEAM ENGINEERING 



Horse Power Constants — Continued 



Piston speed per minute. 


i 


400 


500 


600 


700 


800 


900 


47 


21.030 


26.287 


31.545 


36.802 


42.059147.317 


48 


21.934 


27.418 


32.901 


38.385 


43.868 49.352 


49 


22.858 


28.572 


34.286 


40.001 


45.715'51.429 


.50 


23.800 


29.750 


35.700 


41.650 


47.600 


53.550 


,51 


24.762 


30.952 


37.142 


43.333 


49.523 


55.713 


52 


25.742 


32.178 


38-. 613 


45.049 


51.484 


57.920 


.53 


26.742 


33.427 


40.113 


46.798 


53.483 


60.169 


,54 


27.760 


34.700 


41.640 


48.581 


55.521 


62.461 


,5.5 


28.798 


35.998 


43.197 


50.397 


57.596 


64.796 


,56 


29.855 


37.318 


44.782 


52.246 


59.709 


67.173 


.57 


30.930 


38.663 


46.396 


54.128 


61.861 


69.594 


,58 


32.025 


40.032 


48.038 


56.044 


64.051 


72.057 


,59 


33.139 


41.424 


49.709 


57.993 


66.278 


74.563 


60 


34.272 


42.840 


51.408 59.976 


68.544 


77.112 



FLY WHEELS 

When the speed of an engine is to be 
determined there are several points that 
ought to be considered, one of which is 
the safe speed for the fly wheel. This 
cannot be stated in revolutions per min- 
ute alone, but must be taken in connec- 
tion with the diameter of the wheel. 
For illustration of this point we cannot 
say that a fly wheel should never be run 
more than 300 revolutions per minute, 
because the safe limit for a large wheel 
is much less and for a small one it may 
be a great deal more. 

The safe limit of any wheel is the num- 
ber of feet that the rim or the face of the 



STEAM ENGINES 287 

wheel can travel in a minute without 
danger of disruption by centrifugal 
force. 

The rim speed of a fly wheel may be 
determined by the following rule. 

Multiply the diameter in feet by 
3.1416 and by the number of revolutions 
per minute. The product is the speed 
in feet per minute. 

Now the safe speed of cast iron fly 
wheels was formerly taken at 5,000 feet 
per minute (which is about one mile), 
but as higher speeds are imperatively 
demanded in modern practice, the limit 
has been raised to 100 feet per second or 
6,000 feet per minute. Even this allows 
Si large factor of safety, provided the iron 
is of fair quality, and the casting is free 
from defects, but these desirable quali- 
ties cannot be guaranteed in every case. 
All the parts of a fly wheel should be 
designed with a large factor of safety to 
iwithstand even greater strain than it is 
subjected to under ordinary working 
iconditions, especially in view of the fact 
that if a governor does not control the 
■speed perfectly, allowing it to increase 
above normal conditions, the strain may 
[be greatly increased before the vspeed 
can be reduced. 



288 STEAM ENGINEERING 

When the rim of a fly wheel is made 
thicker than usual it is not a guarantee 
of greater safety, because centrifugal 
force increases with the weight, hence 
the strain is much greater on a thick rim 
than on a thin one, when both are run at 
the same speed, and the danger of hid- 
den flaws in cast iron increases with the 
size of the casting, therefore after a cer- 
tain thickness is secured, based on the 
shape of rim, number of arms supporting 
the same, etc., it is useless to make it 
thicker. , 

The number of revolutions per minute 
that can safely be allowed for well de- 
signed cast iron wheels, free from defec- 
tive castings, is determined by the fol- 
lowing rule: 

Divide 6,000 by the diameter multi- 
plied by 3.1416. The quotient is the 
safe number of revolutions per minute. 
For illustration take a wheel 10 feet in 
diameter. Then 6,000-j- (10X3.1416) = 
191 revolutions. 

The next table gives the safe speed of 
wheels from 4 to 30 feet in diameter 
based on a rim speed of 6,000 feet per 
minute. 

For well designed and constructed 
wooden fly wheels, it is probably safe to 
add 25 per cent, to the figures given in 
the table. 



STEAM ENGINES 289 

SAFE SPEED OF FLY WHEELS 



Diameter 


Revolutions 


Diameter 


Revolutions 


in feet 


per minute 


in feet 


per minute 


3 


636 


17 


112 


4 


477 


18 


106 


6 


381 


19 


100 


6 


318 


20 


95 


7 


272 


21 


91 


8 


-238 


22 


86 


9 


212 


23 


83 


10 


191 


24 


79 


11 


173 


25 


76 


12 


159 


26 


73 


13 


146 


27 


70 


14 


136 


28 


68 


15 


127 


29 


66 


16 


119 


30 


63 



MORE ABOUT HORSE POWER 

Men who are not engineers and who 
have not given the matter much atten- 
tion^ seem to think that if an engineer is 
told the horse power of an engine that 
he has never seen, he should know at 
once the size of it, or in other words the 
diameter of cyHnder and length of stroke. 
It is impossible to do this because a 
given horse power may be secured from 
a great variety of sizes, and on the other 
hand a given size may represent a wide 
range of power developed under differ- 
ent conditions. 

In order to fully illustrate this matter 
several tables will be given which repre- 
sent good practice at the present time 



290 STEAM ENGINEERING 

as they are published by up-to-date 
engine builders as representing their 
product to prospective customers. For 
the benefit of the reader who is seeking 
practical information along this line, 
they will prove valuable as showing 
combinations of sizes that will give sat- 
isfaction in service. 

The first of these tables shows the wide 
range of power that can be obtained from 
each engine of given size, beginning 
with a small one with a 10X24 inch cyl- 
inder rated at 38 horse power, and end- 
ing with a 34X60 inch cylinder rated at 
647 horse power. There are four ratings 
given for each engine although only one 
speed is included in each case, and the , 
boiler pressure is 80 pounds by the gauge, 
in all cases mentioned in the table, the 
columns of which contain the following 
information : 

D = Diameter of cyhnder in inches. 

S = Stroke in inches. 

R = Revolutions per minute. 

P = Piston speed in feet per minute. 

V= Variation in power without seri- 
ously affecting economy. 

M = Maximum power that can be 
secured under given condi- 
tions. 

B = Power developed with best 
economy. 







STEAM 


ENGINES 




291 


HORSE POWER WITH 80 LBS. 


PRESSURE 


D 


S 


R 


P 


V 


M 


B 


10 


24 


110 


440 


33 to 43 


56 


38 


12 


28 


95 


443 


48 " 


59 


76 


53 


14 


32 


85 


453 


68 ' 


83 


109 


75 


16 


30 


90 


450 


90 ' 


109 


142 


100 


16 


38 


85 


538 


106 ' 


128 


170 


117 


16 


42 


85 


595 


119 ' 


144 


188 


131 


18 


36 


90 


540 


135 ' 


151 


213 


143 


18 


42 


85 


595 


149 ' 


181 


234 


165 


20 


42 


80 


560 


174 ' 


211 


275 


192 


20 


48 


75 


600 


186 ' 


227 


300 


205 


22 


42 


75 


525 


197 ' 


240 


313 


218 


22 


48 


75 


600 


225 * 


274 


360 


250 


22 


52 


75 


649 


244 • 


296 


390 


270 


24 


42 


82 


560 


257 ' 


312 


398 


284 


24 


48 


75 


600 


270 ' 


326 


429 


295 


24 


56 


70 


653 


294 ' 


350 


465 


322 


26 


48 


75 


600 


315 • 


383 


503 


350 


26 


54 


70 


630 


331 • 


400 


528 


367 


26 


60 


65 


660 


344 • 


416 


545 


380 


28 


54 


67 


603 


367 • 


444 


584 


405 


28 


60 


65 


650 


400 ' 


483 


632 


440 


30 


60 


65 


650 


459 • 


556 


726 


507 


32 


48 


80 


640 


510 * 


619 


809 


564 


32 


60 


65 


650 


520 ' 


633 


826 


576 


34 


60 


65 


650 


585 ' 


710 


932 


647 



The indicated horse power of an en- 
gine varies directly with the speed, pro- 
vided this is not excessive, for it was 
probably designed to run at a given rate 
and if the number of revolutions is in- 
creased, the mean effective pressure may 
not be as high as it was with a lower 
speed, although the boiler pressure re- 
mains unchanged. While the preceding 
table gives various ratings at SO pounds 
boiler pressure, the next states the indi- 
cated horse power that can be realized 
with higher boiler pressures. Both 
relate to simple, non-condensing engines 



292 



STEAM ENGINEERING 



The several colurflns contain the follow- 
ing information. 

D= Diameter of cylinder in inches 

S = Stroke in inches. 

R = Revolutions per minute. 
There are two columns under 90 
pounds, the first giving the builder^s 
rating when steam is cut off at Vs stroke, 
and the second when the point of cut off 
is lengthened to j^ stroke, which is not 
unreasonable. The next two columns 
give the ratings with the same points of 
cut off, provided the boiler pressure is 
increased to 100 pounds, while the next 
two give ratings based on the same points 
of cut off, with a boiler pressure of 110 
pounds by the gauge. 



INDICATED HORSE POWER— LOW SPEED 





S 


R 


90 


100 


110 


D 


\k 


H 


\% 


H 


hi 


H 


11 


24 


110 


50 


79 


55 


65 


61 


72 


12 


30 


90 


62 


54 


69 


83 


77 


92 


12 


36 


85 


70 


84 


78 


94 


87 


105 


14 


36 


85 


95 


114 


107 


128 


121 


142 


16 


36 


82 


120 


144 


135 


162 


151 


180 


16 


42 


78 


133 


159 


150 


179 


168 


200 


18 


36 


80 


148 


177 


166 


199 


186 


222 


18 


42 


78 


168 


202 


189 


227 


212 


253 


18 


48 


75 


195 


222 


208 


249 


234 


271 


20 


42 


75 


200 


240 


225 


270 


252 


300 


20 


48 


72 


219 


263 


246 


296 


275 


330 


22 


42 


75 


242 


290 


271 


326 


303 


364 


22 


48 


72 


265 


318 


298 


358 


333 


400 


24 


48 


70 


307 


368 


345 


414 


386 


460 


26 


48 


70 


360 


432 


405 


486 


454 


541 


28 


48 


68 


406 


487 


457 


548 


514 


595 


30 


48 


68 


444 


526 


507 


694 


590 


683 



STEAM ENGINES 293 

LOW AND HIGH SPEED 

When the speed of an engine is men- 
tioned, it usually refers to revolutions 
of the crank shaft, but it may mean the 
piston speed in feet per minute. How- 
ever, the latter may be the same for two 
engines where the former differs widely. 
For illustration, take a 12x36 inch en- 
gine at 90 revolutions, giving a piston 
speed of 540 per minute. If a 12X12 
inch engine revolves three times as fast, 
making 270 revolutions, the piston 
speed will be 540 feet as in the preced- 
ing case, but the former is generally 
called a low speed engine, while the latter 
is known as a high speed machine. 
With the same initial pressure, and cut- 
ting off at equal points in the stroke, the 
indicated horse power is alike, therefore 
the difference between them is not so 
great as it appears at first. 

If power is to be transmitted by belt 
from the engine to a shaft in a mill, shop 
or factory, the low speed engine with a 
long stroke is appropriate, because it 
will prove much more durable owing to 
the fact that the valve gear reverses its 
motion only 180 times per minute, and it 
is the reversing process that causes wear 
and tear in an engine. If a comparative 
high speed must be obtained at once, 
making it necessary to connect the en- 



294 STEAM ENGINEERING 

gine directly to the machine to be driven 
as in the case of a dynamo or generator, 
the high speed engine with a short stroke 
must be used, but the valve gear is of a 
very different type, as a general rule, for 
its direction of travel must be reversed 
540 times per minute, therefore, any 
form of cut off in which a moving part 
engages a part that is at rest (as with 
the Corliss valve gear) is impracticable 
and cannot be used successfully. 

The following table gives the indicated 
horse power of high speed engines under 
various conditions. The sizes given 
refer to the diameter of cylinder and 
length of stroke. The speed mentioned 
is the number of revolutions per minute. 
Three rates are given for each engine, 
showing the range of speed for which it 
is adapted. Boiler pressures from 80 to 
120 pounds are included with cut off at 
3^, H and }4 stroke. The former should 
not be exceeded in general service, but 
if a heavy load is thrown on occasion- 
ally, the longer points of cut off may be 
utilized for the emergency. When 
selecting an engine of this class it should 
always be large enough to carry the esti- 
mated load, when running at the lowest 
speed given, and cutting off steam at the 
shortest point mentioned. 



STEAM ENGINES 



295 



1 

o 


§ 


05.-|^occo(Nooococ5l^o^•lOO^o 


o 




8 


Ot>-00M«-'5t0'-iC0C0O-*00r000(M 


§ 


rf<iCir)OlMC000O(NCD05C<100(MtO 
r-iT-«i-lC<)(N(NlMCCCOCOCO->#'*iCiO 


§ 




1 

a 

fa 
o 

1 


1 


t^ 00 era "O t>. 00 CO CD CT) Tj< t^ O 00 crj t^ 

r-lrH,-^(N(MIMOOCOMTfl'J<lOlOCOO 


o 


sss§^ssg?:?i?§^^SJ;53 


I 


;2:22g§^g5^^^^S^^i:5iS 


§ 


(Nf0-*00C2-HTf-<O00(M>-hI>C^CDCn 


§ 


-4(NfOCOt^OO-<fOL':)000(Mt^OCO 


1 

O 
O 


o 


Tt* lO O C IM -# 00 O <M O 05 <M 00 (M CO 
rH .-1 ^ (M (M IM (M OO M CO CO ■* r}< lO IC 


o 


coTfiooo-'iot^Ofocooo^r^i-H 

r-i,-i,-l^<MCvI(MCvIlMC0fOCO-*'-*iO 


1 


.-" (N CO t^ 00 cr- CO lO CD o 03 -* o CO to 

,_(rHi-li-i^,-i(M(NIMC0COC0C0'4*'<i< 


§ 


O— '<Mi0cDr^O(MC0CD00OL':0C'^ 

ii^^^^^o^(^,(N(M<Mco^ocor}^ 


§ 


ooooeoTi'irooooi'-'co'Ooo-'coto 

^^^^^^C^C<1(M(MCOCOCO 




cocococococococococococococ-oco 




2 


o 
io''co'"i^''oo"'oo'- 



296 



STEAM ENGINEERING 



1 
1 


§ 


«^^2§S§SS23Sg§S 


O 


i0r-if,e0O00(N'HOC0C<J^N-^»0 
t^ 00 00 05 O O »H (M CO OJ 1-1 r-l W CO 


8 


<©t^t^00 05 0t-ttHOOO>00'HN 


§ 


CO«Ot^l^OOOOOiO>Ot>.OOOOOSOO 


§ 


SfoSSnS^JSSSS^J^JSoS 


.•8 


1 


?2^§§g§|2g88§||g 


o 


CO C<J 00 (N OJ CO 05 00 «0 (N <-t OS O O OJ 
«D t^ t^ 00 00 OS OS O >-( 00 OS 05 O !-< ^ 


8 


gsg?:§sss;|?:sS8g| 


§ 


SgS§^?:§§-g§f2?^S§8§ 


§ 


t^ r-l lO 00 CO 00 O CO (M 00 ■* O O 00 lO 
Tl* lO »0 <0 CO CD t^ t>. 00 «5 CO tN. t>. t^ 00 


:5 

1 


S 


S CO t^ t^ 00 00 OS OS OX^ 00 CJ> OS o o 


o 


U5 CO CO CO t>. 00 00 00 OS CO t^ 00 00 OS OS 


I 


g;s§S5?2^§goSS^^g§ 


§ 


Tl<^U51OlOCOCOI>t>.lOCDCOCOt^00 


S 


OS N lO 00 (N CO 00 CO 00 00 CO 00 00 rj< O 

co<<^nH>*ioioiococD'*ic»»oiocot>. 


OT 


OiOOOiOOOiOOOLOOOiOO 
OlMiOO<MiOO(NiOiOt^O»Ot>0 
CO CO CO CO CO CO CO CO CO (N (M CO <N (N 00 




1 


9x10 
10x10 
11x10 
10x12 
11x12 



STEAM ENGINES 



297 





r-1 


r-lr-.rHr-4,-ICQr-lrH,-lr-llMC<ICSllM(N 


O 


CO-<i<COiOr>-00?OiOt>-«300005'-iCO 


8 


'-<C0iO(NtDOcDO'^00'*O(M»-iO 


§ 


t^coa>^air-iiMiot>.>-nO'-icoor^ 

O•-^(N(NC0lO'-^(^^C0(^0T}^CDl0r>•00 


§ 


^23HBS^S§2§S3SS 


1 
1 


i 




o 


i-l 1-1 i-H rH T-l r-l rHT-li-li-li-trHrHr-tC^ 


8 


sSS§2S3S§SSI§l§ 


§ 


^g2^g5i^8:::?5^g5^:J3S 




S 


•<*! <N .-t OO 00 00 00 1>. CO CO T»< »o 0>(N o 
000>00>0-iOOOSOOr-((N'-tcOTl< 


O 
3 

o 


1-1 


sSSBSSSSsSSssSS 


o 


^sSSSSsSSSSSSSS 


8 


0500t-'^ioiOMrO'*05<-icccookO 

00 05 O O rH (M 03 O ^ O !M coo? '^ lO 


o 


a3t^iOO:|iM^(M.-^^t^0000<Nri<r^ 

t>ooa3G>0'-iooa5oaiO'-H'-t(Nco 


§ 


OJcOfOr-^aJt^iMooicio^oooio 

<C t- 00 00 00 03 1^ OC 00 00 O O 05 O (M 


1 


O lO O O lO O lO O lO lO O »0 lO O lO 






(M (N •^ 'I* '^ 

xc; i<::: «:: «:: «:i 

C^ CO W CO '^ 



298 



STEAM ENGINEERING 



1 

Is 

o 

o 




OCOC-TroOCOCqcO-^OOCNlt^CCXMrH 
C^C^<NC<I(MlM(M(NCO(M0OC<:iCOC'0CO 


O 


COCOI>'^OOcO'-IO^-C>OTt<COiC'-tI> 
N(M(NrHeO(M<M(N<N(M(NcrD(MCOCO 


s 


t^05'-liOI>050lOOCCH>CDt^OC<l 
O ^ Tt< t> 0> rH CT) <M UO (M lO 00 lO O (N 
TH(N<Nr-lrHIMr-l<N (M <N (M (M <M (M CO 


§ 


lO lO lO <X) CD lO 00 O CO CO a> •* Oi 00 1> 
t^05'-not>0}t^O(MOCSlu:)(M»000 


§ 


T}<,-IOOt>^rHaJCDCOOi.-HO-JrHI>(M 

lot^oocoior^iot^osb-oc^ocqio 


eS 

o 
o 


1 


<r0CDOOC0CDi0c0Ol>000000C000 
i-HC0C005'-IC0'-HTl^t>.Tt<l>OI>'-<r}< 
W(N(M>-i(M(N(M(NlM(MC<JCO<MCOCO 


o 


Ti<lOCDCO^lOCO^CDu:)CO.HCO»Ot^ 
O5^COt>O3r-<O5(Mr^<NlO00lO00.-l 
.-H <N N ^ -H (N r-i(N (N (N (N (M (N (M CO 


I 


lO ■* CO CO lO -^ |> 05 CO CO 00 CO 00 t>. o 
t^ Oi -H lO t^ 05 1^ cr> CO O (M lO (N lO 00 « 

rtrtC<,^,H,-(,-(,-((Nc<icqcqc^cvico 


§ 


,_,^^r-<THr-l,-lr-(,-(rHC>JCOCOCO(N 


§ 


t*e<jt*cot^c<ioocc)'<i'Oi0505Cjic^Ti< 
co>ccoc<ico'oco'or>.ioi>05t^oco 


1 

eS 

o 

1 


§ 




o 


rH05t>.C0'-H0SC0Tt<-*r^OC0OI>C0 


1 


rHr-Hi-li-li-liHi-HT-lrHiHi-ICO'-ICOC^) 


2 


05 CO t* »0 C75 CO 1> CO Oi 00 1^ Oi o ^ 
CvJ'*lO>-IC1-^cO-*fOTj(CO00COCT>i-l 


S 


S§§2SSS§SS§s§Ss 




lOOiOOiOOOiOOOiOOOiOO 




1 


Ttt «o CO to eo 
k: : x: : x: : x; : x: ; 

iO -^ >0 CO !> 



STEAM ENGINES 



299 



o 


1 


Tl< CO (M C3J M t-- tC lO lO <X) f-< CD O •* t^ 
t^T-HioOiOOJ'^05-*00-<l<05iOCT>fO 
<NMOOOOCOCOfOCOTi<TjH^-*COirO-<4< 


o 


250 
285 
320 
282 
322 
362 
316 
360 
406 
352 
402 
452 
319- 
359 
399 


s 


CD r^. O Tfi O I> lO lO CD 00 (M t^ 00 rt< O 
NNMCqC^COCgcOCOCOfOTTiMCOCO 


g 


(M(NC<JiM(M(N<N(N(rO<MCOCO<N(NCO 


§ 


CD'-<I>a5l>.CDOOTjHCDOO^a3CD-*(M 
t^O(M03(NiOlMiOOO'*00^(MiOOO 
rH(N(M-<(N(M(M(N(M(M<MCO(N(N(N 


o 
3 

o 


§ 


rJ<00(MiOfO<MOO^iOeO(MO'-l005 

'*HI>r-H>r-HOOlOO)'*CT>Tt<^lOOO 

N cq CO (N CO CO CO CO (^^ CO CO •* CO CO CO 


o 

s 


<NCOiOOiO'-tOOO(Nt>->-i-*00'# 
(N>O00iO00(N00(MCD-^if3O00^»O 
(N(N<N(NiNCOlMrOCOCOCO-*(NCOCO 


O00t^iOt^O(Nb0i0^(N(NCDt^03 
§(MC<liN(M(M<N<NC0(M00CO<N(MC0 


§ 


.-HCSI(NCQ(N<N(M(N<M(NiMCO(N<NiM 


s 


CD(35rH00C0t^00CD-*i-l(Mc0O»OO 


Is 
o 

"3 
o 


§ 


OC0CD(MCD05i0CT>(M00iMCDiO05(N 
(N(M<Nl?a(N<N(MiMCO<NC0CO(NC^C0 


o 


t-i(MCNJ(NC<J<N<M(MCO(MiMC0<M(N(N 


8 


«o o •* r^ •«* r-t o o o Tf< f^ o CO OS »o 


§ 




§ 


0500C0COr-C0TlHt^O(N00Tt<iCcDt>. 


1 


lO O lO lO O lO lO O lO »0 O lO o o o 




en 


00 00 "^ °o <=> 
'><::'x;:'x::'«::*»<:: 

CO t* 00 OJ « 



300 



STEAM ENGINEERING 






Pi 
o 

K 

Q 
» 

i 

Q 
IS 



1 

la 
P 


o 


OOiOOTtHt^i-it^t^t^t^NtOMOO 
05COOOCOCO-*t^0005'*03COO>'<*0> 


O 


to O lO to -* CO lO O T}< 00 00 00 00 CO t* 
»0 O ■* 0> -^ 05 CO 05 Tj< o •<* 00 -"l* 0> CO 
00 ■<1< ■* CO ^l* 'i^ ■* ■* »o ■<*< Tj< •* ■* TJHO 


8 


rHrHi-ir^.-llO(MIN'-IOO»Or-tTt<lOlO 
CO to O lO O ■* 03 ■* CT> to O TJH O Tt< 00 
C0CO-<*00-*-*CO-*'*00-<l<Tj<Tl<-^Ti( 


8 


CO CO CO CO CO CO CO CO ^ CO COCO CO CO^^ 


§ 


C0C0'^OS'*00t^tOTj<00JN«OCO00O 

c«c<icococooococococaoocooococo 


1 


i 


CO o CO ko CO o ■>* to o 00 r* to to Q ''Jt 


o 


to U5 ■* O ■* t^ CO •* CO CO 00 ■<* t^ t* t>i 


8 


00 N lO tH lO CS rj< 0> CO CO lO O lO 0> 00 

cocoeococooococo'<J<cococococo^ 


8 


•>!}<tO00C0t^C0i-HOO5C0OC35OC0Tj< 

»ooO'-ioo.-i>o.-(ioooo5cq-<*icokooo 
c^cooococooocooococococoooooco 


8 


00 rH c:> 00 00 0> CO t^ r-l to 1-1 1^ .-( o> t^ 


5 

o 


§ 


00 -^ O ^ Q O CO N. CO 1-1 ■* t^ CO 00 lO 


o 


CO U5 00 CO 00 •«* O rH CO .-1 1-1 ,-H rH rj( to 


8 


t* CO to 00 to 00 c:5 to CO .-1 00 >o 00 00 00 

00 to C» to 05 CO 00 CO to t^ 05 CO 05 CO >o 

cqcqcococoooc<icocoeocococococo 


o 

05 


rH I> CO •* CO CO t^ CT> CO .-llO OS lO CO 00 
i-HCOCOC0tDO3lO00C^-<*tO00tOC3Si-l 
CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO 


g 


-d<00rHiOOtOtDTt<C0.-<C0TtlC0tO00 
OOOCOOCOiOCOiOOO'-HCO"3eOU5t>. 
T-tCOCOCOCOCOCOCOCOCOCOCOCOCON 




§|8|g|||8|||||| 




(0 


19x20 
20x20 
21x20 
20x22 
21x22 



STEAM-ENGINES 



301 



INDICATED HORSE POWER— HIGH 
SPEED 





Speed 


Cut off at H stroke 


Size 


80 


90 


100 


110 


120 


22x22 


150 


255 


291 


328 


363 


399 




165 


281 


320 


361 


400 


439 


" 


IHO 


30« 


349 


393 


436 


479 


23x22 


150 


280 


319 


359 


398 


437 




165 


308 


351 


395 


438 


481 




180 


336 


383 


431 


478 


525 





Speed 


Cut off at H stroke 


Size 


80 


90 


100 


110 


120 


22x22 


150 


309 


352 


394 


437 


480 


" 


165 


340 


387 


434 


481 


527 


•' 


180 


371 


422 


473 


524 


575 


23x22 


150 


338 


386 


433 


479 


526 


"• 


165 


372 


424 


475 


527 


578 




180 


406 


462 


518 


574 


630 





Speed 


Cut off at H stroke 


Size 


80 


90 


100 


110 


120 


22x22 


150 


348 


396 


444 


492 


540 


'* 


165 


383 


435 


488 


542 


594 


*' 


180 


418 


475 


533 


591 


648 


23x22 


150 


381 


434 


487 


540 


592 




165 


420 


478 


536 


594 


652 




180 


457 


521 


585 


648 


711 



^ 



302 STEAM ENGINEERING 



COMPOUND ENGINES 



Many steam engineers seem to believe 
that the principal object in adopting a 
compound engine is to secure more 
power than could be developed in any 
other kind occupying the same space. 
This idea is probably based on the well 
known fact that two cylinders are used 
and as they are supposed to develop 
more power than could be secured from 
one, it constitutes conclusive proof in 
their estimation. It follows as a nat- 
ural consequence that if pressure in the 
receiver between the two cylinders is 
light, the second cylinder is considered 
of little or no value, because, it is 
claimed that the small piston does all 
of the work, and in addition to this, it 
must drive the large piston. 

It is possible to find a few places where 
this state of affairs exists, but they are 
not so common as engineers who are not 
well educated along this line seem to 
beheve. One reason for this belief is 
found in the following statement of facts: 
A small piston driven by high steam 
pressure may operate against a com- 
paratively high back pressure, therefore, 
the net power available for driving ma- 
chinery is not as great as the initial pres- 
sure indicates when considered alone. 
On the other hand a large piston may be 



( 



STEAM ENGINES 303 

driven by a low steam pressure, but it 
operates against a very low back pres- 
sure, consequently the net power avail- 
able for useful work is nearly or quite 
equal to that secured in the high pres- 
sure cylinder. 

If a simple engine is overloaded until 
it cannot maintain the required speed, 
when a superior cylinder oil is used to 
lubricate its internal parts, and the 
valves are properly set, the best way to 
secure more power is to remove the old 
machine and put in another simple en- 
gine with a cylinder large enough to do 
the work easily, provided economy in 
the use of steam is not of great impor- 
tance. If a simple engine is run under 
conditions that prove wasteful of steam, 
the remedy is to install a compound 
engine that is well adapted to the neces- 
sary load. Where the exhaust steam 
can be used profitably, it should be run 
non-condensing, but otherwise a con- 
denser should be added to remove back 
pressure from the large piston. 

From this it will be plain that the real 
object in installing a compound engine 
is to develop power with the least pos- 
sible amount of steam. This is consistent 
because this type makes it practicable 
to actually use steam at high pressure 
by expanding it to the greatest profit- 
able extent, and at the same time cylin- 



304 STEAM ENGINEERING 

der condensation is reduced to a lower 
point than can be secured when expand- 
ing steam to a low pressure in a single 
cylinder. 

If a very high pressure is used in a 
simple engine carrying a light load, the 
terminal pressure will be low, and the 
temperature of the cylinder at the end of 
each stroke will be low, consequently 
when another charge of steam is admit- 
ted, some of it will be condensed in 
raising the temperature of the cylinder, 
thus causing a loss of heat. With a 
compound engine this difference is 
divided between two cylinders, hence it 
is less for each, resulting in economical 
use of the steam. 

DESIGNING COMPOUND ENGINES 

A very good method of designing a 
simple engine, assuming that 300 horse 
power would be required, has been illus- 
trated on previous pages, and in order 
to show the economy of the compound 
engine it becomes necessary to design 
one for a load of 300 horse power, and 
compare the results. It is assumed 
that this engine is run non-condensing. 

Modern practice with these engines 
has demonstrated that within certain 
reasonable limits the pressure, total ratio 
of expansion, and comparative size of 



STEAM ENGINES 305 

cylinders must follow general rules in 
order to secure good results, and these 
have been observed in the example 
given. Data to be used as a basis for 
the calculations may be stated as follows : 

Gauge pressure at the boiler, 130 pounds 
Atmospheric pressure, 15 " 
Initial pressure absolute, 140 " 
Ratio of expansion, high 
• pressure cylinder, 3 
Ratio of expansion, low pres- 
sure cylinder, 2.5 
Total ratio of expansion, 7.5 

TO DEVELOP 300 HORSE POWER 

In order to make an intelligent com- 
parison that can be easily imderstood 
and appreciated, the stroke of this engine 
is assumed to be 42 inches and the speed 
90 revolutions, making the piston speed 
630 feet per minute, to correspond with 
the simple engine above mentioned. 

The first point in the problem is to 
determine the mean effective pressure 
that will result from these conditions, 
using the following formula. 



SZPi±S±i22<P_B = MEP. 



306 STEAM ENGINEERING 

Hyp Log = Hyperbolic logarithm of 
the ratio of expansion. 
P = Initial pressure absolute. 
R = Ratio of expansion. 
B = Back pressure absolute. 
M E P = Mean effective pressure. 

Applying it to these conditions gives 
the following results: 

(2.0149 + 1) X 140 ,^ ^, 

15=41 pounds 

7.5 

mean effective pressure. 

When designing a compound engine it 
is assumed that all of the power is de" 
veloped in the low pressure cylinder, 
consequently it becomes necessary to 
determine the diameter of a cylinder 
that will prove large enough for this pur- 
pose, after which it will be used for a low 
pressure cylinder. A high pressure 
cylinder will then be designed and the 
total power developed is to be divided 
between them as nearly even as possible. 
The formula for a simple engine must be 
used here. 

H-PXF-P 
MEPXS" 






STEAM ENGINES 307 

H-P =Horse power required. 
F-P=Foot pounds for 1 horse 
power. 
M E P = Mean effective pressure. 

S= Piston speed in feet per 

minute. 
A = Area of piston. 

Applying it to stated conditions gives 
the following result: 

300X33,000 „„„ . , 

=383 square inches. 

41X630 ^ 

Referring to the table of the area of 
circles we find that the nearest area is 
380 square inches, which corresponds to 
a circle 22 inches in diameter, therefore 
the cylinder is 22X42 inches. Proof is 
shown by calculating the power that 
this engine will develop as follows: 

380X41X630 ^^^ ^ ^ 

=297.4 horse power 

33,000 . 

which is near enough for all practical 
purposes, as this is an estimate only. 

Now the total ratio of expansion in a 
compound engine is found by multiply- 
ing the ratio in the high pressure cylin- 
der by the ratio of the two cylinders, or 
by the comparative areas of them. To 
illustrate this point we may assume that 



308 STEAM ENGINEERING 



1 



the ratio of expansion in the high pres- 
sure cylinder is 3. The area of the low 
pressure cylinder is 2.5 times as great 
as the high pressure, therefore the total 
ratio of expansion is 3X2.5=7.5. The 
stroke of the two cylinders is supposed 
to be eq ^al in these explanations. 

Taking these facts into consideration, 
and having the area of the low pressure 
cylinder given, by dividing it by the 
ratio between the two, the quotient 
determines the area of the high pressure 
cylinder. In this case it is 380-^2.5 = 
152 square inches, which nearly corre- 
sponds to the area of a circle 14 inches in 
diameter, therefore the high pressure 
cylinder is 14 inches, with an area of 
153.93 square inches. The mean effective 
pressure for the low pressure cylinder 
alone is 41 pounds, but when considering 
the complete engine, one-half of this 
must be applied to the low pressure 
cylinder, or 41^2 = 20.5 pounds. When 
the other half is multiplied by the ratio 
between the two cylinders, the product 
is the necessary mean effective pressure 
for the high pressure cylinder, and 20.5 
X 2. 5 =51.25 pounds. This process is 
known as raising the mean effective 
pressure of the low pressure cylinder to 
terms of the high pressure. Under 
these conditions this cylinder will 
develop — 



STEAM ENGINES 309 

153.93X51.25X630 



33,000 
power. 



= 150.6 horse 



The low pressure cylinder will de- 
velop — 

380X20.5X630 ,,„ ^^ 

= 148 . 7 horse power. 

33,000 ^ 

The whole engine will develop 150.6 
+ 148.7=299.3 horse power. 

THE MEAN EFFECTIVE PRESSURES IN 
COMPOUND ENGINES 

It now becomes necessary to deter- 
mine the mean effective pressures that 
can be secured under given conditions, 
and compare them with those required 
to secure the desired power, as before 
demonstrated. Before proceeding to do 
this a few general rules must be consid- 
ered in order to make the matter plain. 

1. When the initial pressure in the high 
pressure cylinder of a compound engine 
is divided by the ratio of expansion for 
that cylinder, the quotient is the termi- 
nal pressure. 

2. The terminal pressure in the high 
pressure cylinder is equal to the back 
pressure in the same cylinder. 



310 STEAM ENGINEERING 

3. The back pressure in the high pres- 
sure cylinder is equal to the initial pres- 
sure in the low pressure cylinder. 

4. When the initial pressure in the low 
pressure cylinder is divided by the ratio 
between the two cylinders, the quotient 
is the terminal pressure for that cylinder 
and for the whole engine. 

5. The terminal pressure in the low 
pressure cylinder bears no definite rela- 
tion to the back pressure in it, but they 
should be nearly equal in the case of a 
non-condensing engine. 

6. The ratio of expansion for the low 
pressure cylinder is fixed by the compar- 
ative areas, or the ratio between the two 
cylinders. Thus, if the ratio of cylinders 
is 1 to 4 the ratio of expansion for the 
low pressure is 4 regardless of other 
conditions. 

7. The total ratio of expansion divid- 
ed by the ratio of cylinders, gives the 
ration of expansion for the high pressure 
cylinder. 

In this case the initial pressure is 140 
pounds, and the ratio of expansion is 3, 
therefore the terminal and the back 
pressure is 140-7-3 = 46.66 pounds. The 
mean effective pressure for this cylinder 



STEAM ENGINES 311 

(1.0986+1) X 140 ^^^^ ^^ ^^ 

^^ —^ 46.66=51.27 pounds. 

3 

The mean effective pressure for the 
low pressure cylinder is — 

(.9163 + 1) X 46.66 ,^ ^^^^ 
15 = 20.76 pounds. 

These pressures very nearly agree 
with those determined by a different 
calculation, which proves their value. 

The total power of this engine using 
the pressures found by the latter calcu- 
lation is next given. For the high pres- 
sure cylinder it is — 

153.93X51.27X630 _^ ^ , 

= 150,6 horse power. 

33,000 

. For the low pressure cylinder it is — 

380X20.76X630 



33,000 



= 150.6 horse power. 



The whole engine develops 150.6+ 
150.6 = 301.2 horse power. 

The next table contains a list of long 
stroke compound non-condensing en- 
gines that are proportioned to suit the 
given steam pressure by the gauge, and 
when developing the power stated they 
will prove economical in the use of steam. 



312 



STEAM ENGINEERING 



They will develop 25 per cent, more 
power without seriously affecting their 
efficiency, and the speed may be in- 
creased 20 per cent, with a correspond- 
ing increase in available power. How- 
ever, the speeds given in the table are 
about right in order to secure durability 
of the valve gear and noiseless operation 
of the same. 

LOW SPEED COMPOUND NON-CON- 
DENSING ENGINES. 
For 100 pounds gauge pressure. 



Size 


1 


t 


Size 


1 


1 




CO 


(2 




w 


A^ 


ll&16x30 


100 


100 


26&38x60 


65 


734 


ll&16x36 


90 


108 


27&40x54 


70 


782 


12&18X36 


90 


135 


27&40x60 


65 


812 


12&18x42 


80 


140 


28<fe42x54 


70 


8f69 


14&20x30 


100 


155 


28&42x60 


65 


897 


14&20x36 


90 


167 


30&44x54 


70 


968 


15&22x36 


90 


205 


30&44x60 


65 


988 


15&22x44 


80 


213 


31&46x66. 


60 


1,095 


16&24x36 


90 


243 


31&46x72 


55 


1,095 


16&24X42 


80 


252 


32&48x66- 


55 


1,089 


18&26X42 


80 


297 


32&48x72 


50 


1,090 


18&26x48 


75 


318 


34&50x66 


55 


1,286 


19&28x42 


80 


342 


34&50x72 


50 


1,176 


19&28x48 


75 


366 


35&52x56 


55 


1,283 


20&30X42 


80 


392 


35&52x72 


50 


1,272 


20&30X48 


75 


420 


36&54x66 


55 


1,385 


22&32x48 


75 


480 


36&54x72 


60 


1,374 


22&32x54 


70 


504 


38&56x66 


55 


1,488 


23&34x48 


75 


546 


38&56x72 


50 


1,47& 


23&34x54 


70 


573 


39&58x66 


55 


1,697 


24&36x54 


70 


636 


39&58x72 


50 


1,684 


24&36x60 


65 


656 


42&62x66 


65 


1,821 


26&38x54 


70 


712 


42&62x72 


50 


1,806 



STEAM ENGINES 



313 



LOW SPEED COMPOUND NON-CON- 
DENSING ENGINES. 

For 125 pounds gauge pressure. 



Size 


13 

1 


1 


Size 


a 


1 




03 


Ph 




t/j 


^ 


9&14x30 


inn 


92 


22&34x54 


70 


687 


9&14x36 


on 


99 


22&34x60 


65 


708 


10&16x36 


9n 


130 


23&36x54 


70 


769 


10&16X42 


8n 


135 


23&36x60 


65 


793 


11&18X30 


100 


150 


24&38x54 


70 


857 


ll«S:18x36 


9n 


162 


24&38x60 


65 


884 


13&20X36 


90 


203 


25&40x66 


60 


900 


13&20X42 


80 


210 


25&40x72 


65 


990 


14&22x36 


9a 


246 


26&42x66 


55 


1,004 


14&22x42 


80 


255 


26&42x72 


50 


996 


16&24X42 


80 


302 


28&44x66 


55 


1,183 


16&24X48 


75 


324 


28&44x72 


50 


1,092 


17&26x48 


75 


357 


29&46x66 


55 


1,204 


17&26x60 


65 


382 


29&46x72 


50 


1,194 


18&28X42 


80 


413 


30&48x66 


55 


1,313 


18&28X48 


75 


443 


30&48x72 


50 


1,302 


I9&30X48 


75 


509 


32&50x66 


55 


1,422 


19&30X54 


7(1 


534 


32&50x72 


50 


1,410 


20&32x54 


7r 


607 


34&54x66 


55 


1,658 


20&32X60 


65 


627 


34&54x72 


50 


1,644 



LOW SPEED COMPOUND NON-CON- 
DENSING ENGINES 

For 150 pounds gauge pressure 





v. 


S 




1 


u 


Size 


s 


o 


Size 


% 




w 


cu 




w 


CU 


9&14x30 


mo 


100 


14&22x36 


90 


267 


9&14x36 


90 


108 


14&22x42 


80 


277 


10&16X36 


90 


141 


16&24x42 


80 


329 


10&16x42 


MO 


146 


16&24X48 


75 


353 


ll&lSxSO 


TOO 


165 


17&26X42 


80 


38& 


ll&18x36 


90 


179 


17&26X48 


75 


414 


13&20x36 


90 


220 


18&28x42 


80 


448 


13&20X42 


80 


228 


18&28X48 


75 


480 



314 



STEAM ENGINEERING 



LfOW Speed Compound Non-Condensing Engines 51 
Continued 



Fo 


r 150 


pounds 


gauge pressure 






"S 


s 




?, 


S 


Size 


s. 


1 


Size 








w 


Ph 




w 


CIh 


19&30x48 


75 


551 


26&42x66 


55 


1,089 


19&30x54 


70 


579 


26&42x72 


60 


1,080 


20&32x54 


70 


659 


28&44x66 


55 


1,195 


20&32X60 


65 


680 


28&44x72 


50 


1,186 


22&34x54 


70 


743 


29&46x66 


55 


1,307 


22&34x60 


65 


767 


29&46x72 


50 


1,296 


23&36x54 


70 


832 


30&48X66 


55 


1,423 


23&36x60 


65 


868 


30&48x72 


50 


1,400 


24&38x54 


70 


929 


32&50X66 


55 


1,544 


24&38x60 


65 


958 


32«&50x72 


50 


1,532 


25&40x66 


60 


1,076 


34&54x66 


55 


1,801 


25&40x72 


55 


1,076 


34&54x72 


50 


1,786 



HIGH SPEED COMPOUND ENGINES 

The long stroke engine with a com- 
paratively low rotative speed, which 
gives a high piston speed, is generally 
considered the best for shop, mill and 
factory service, where high speed of the 
jack shaft can easily be secured by well 
proportioned pulleys and good belts, but 
there are cases where a high rotative 
speed is wanted without using belts, 
therefore, a high speed engine must be 
used. 

The' next table gives sizes of non- con- 
densing engines that can be used with 
stated pressures to develop the power 



STEAM ENGINES 



315 



{given in each case. As the cylinder 
ratio is 1 to 3.5 the higher pressures can 
be used to good advantage. The first 
column contains the diameter of cylin- 
ders and length of stroke. The second 
is the speed in revolutions per minute. 
Succeeding columns give the power 
developed. 

HIGH SPEED COMPOUND NON-CON- 
DENSING ENGINES 



Cut of! at }i Stroke. 



Size 


1 




Initial 


pressure 
















^ 


100 


110 


120 


130 


140 


150 


7&11X10 


300 


30 


34 


37 


41 


44 


48 


" 


325 


32 


36 


40 


44 


48 


62 


*• 


350 


34 


39 


43 


47 


52 


66 


8&13xl0 


300 


40 


45 


60 


55 


60 


65 




325 


43 


49 


54 


59 


65 


70 


•• 


350 


46 


52 


58 


64 


69 


75 


8&13xl2 


250 


40 


45 


50 


55 


60 


65 




275 


44 


50 


55 


60 


66 


71 




300 


48 


54 


60 


66 


72 


78 


9& 16x12 


250 


55 


62 


68 


75 


82 


89 




275 


60 


68 


75 


83 


90 


98 




300 


66 


74 


82 


90 


99 


107 


9& 16x14 


225 


57 


65 


72 


79 


86 


94 




250 


64 


72 


80 


88 


96 


104 


** 


275 


70 


79 


88 


96 


105 


114 


ll&18xl4 


225 


80 


90 


100 


109 


119 


129 




250 


88 


100 


111 


122 


132 


144 


, " 


275 


97 


110 


122 


133 


145 


158 


ll&18xl6 


200 


81 


91 


101 


111 


121 


131 


'• 


225 


91 


102 


114 


124 


136 


148 




250 


101 


114 


127 


139 


152 


164 



316 



STEAM ENGINEERING 



HIGH SPEED COMPOUND NON-CON 

DENSING ENGINES 

Cut oflE at 14 Stroke — Continued 



Size 



Initial Pressure 



100 110 120 130 140 150 



12&20xl6 
12&20X18 
14&23xl8 
15&25x20 
17&28x20 
17&28x22 
19&32x22 



Cut off at H Stroke. 





-s 


Initial pressure. 


Size 
















w 


100 


110 


120 


130 


140 


150 


7&ll:il0 


300 


41 


45 


50 


55 


60 


65 




32ft 


44 


49 


55 


59 


65 


70 


'♦ 


350 


47 


53 


59 


64 


70 


76 


8&13xl0 


300 


55 


61 


68 


74 


81 


88 




325 


59 


66 


73 


81 


88 


95 


" 


350 


64 


72 


79 


87 


94 


102 


8&13xl2 


250 


55 


61 


68 


74 


81 


88 




275 


60 


67 


75 


82 


89 


97 


" 


300 


66 


74 


81 


89 


97 


105 



k. 



STEAM ENGINES 



317 



HIGH SPEED COMPOUND NON-CON- 
DENSING ENGINES 

Cut off at % Stroke — Continued 





1 


Initial pressure 


Size 
















^ 


100 


110 


120 


130 


140 


150 


9&16xl2 


250 


76 


85 


94 


103 


112 


~122 


" 


275 


83 


93 


103 


114 


124 


134 


•• 


300 


91 


102 


113 


124 


136 


146 


9&16xl4 


225 


79 


89 


99 


108 


118 


128 


" 


250 


88 


99 


110 


121 


131 


142 




275 


97 


109 


121 


132 


144 


156 


9&18xl4 


225 


109 


123 


136 


149 


162 


176 


♦• 


250 


121 


136 


151 


165 


180 


195 


•• 


275 


134 


149 


166 


182 


198 


214 


ll&18xl6 


200 


111 


125 


138 


151 


164 


178 


'< 


225 


125 


410 


155 


170 


185 


201 


" 


250 


139 


156 


712 


189 


206 


223 


12&20X16 


200 


134 


151 


167 


183 


199 


216 


" 


225 


151 


170 


188 


206 


224 


243 


" 


250 


168 


189 


209 


229 


249 


270 


12&20X18 


175 


132 


148 


164 


180 


196 


213 


•• 


200 


151 


170 


188 


206 


224 


243 




225 


170 


191 


211 


232 


252 


274 


14&23xl8 


175 


178 


199 


220 


242 


263 


285 




200 


202 


228 


252 


277 


301 


326 


«• 


225 


224 


256 


284 


311 


338 


367 


15&25x20 


160 


212 


237 


263 


288 


314 


340 


" 


180 


238 


267 


296 


324 


353 


383 


•• 


200 


264 


297 


328 


361 


392 


426 


17&28x20 


160 


266 


299 


331 


364 


395 


429 


" 


180 


300 


336 


372 


409 


445 


482 


" 


200 


333 


374 


414 


454 


494 


536 


17&28x22 


150 


274 


308 


342 


376 


407 


442 




165 


303 


339 


376 


413 


449 


486 


" 


180 


330 


370 


410 


449 


489 


531 


1 19&32x22 


150 


350 


394 


436 


479 


520 


565 


" 


165 


38C 


433 


479 


526 


573 


621 


.. 


180 


420 


472 


522 


574 


624 


678 



318 



STEAM ENGINEERING 



HIGH SPEED COMPOUND NON-CON- 
DENSING ENGINES 

Cut off at 14 Stroke 







Initial pressure 


Size 


^ 
















a 


100 


110 


120 


130 


140 


150 


7&llxl0 


300 


47 


52 


58 


63 


69 


75 


" 


325 


51 


57 


63 


69 


75 


81 


" 


350 


55 


61 


68 


74 


81 


87 


8&13xl0 


300 


63 


71 


79 


86 


94 


101 


" 


325 


69 


77 


85 


93 


101 


110 


" 


350 


74 


83 


92 


100 


109 


118 


8&13xl2 


250 


63 


71 


79 


86 


94 


101 


" 


275 


70 


78 


86 


95 


103 


111 


" 


300 


76 


85 


94 


103 


112 


122 


9&16xl2 


250 


88 


99 


110 


120 


131 


142 




275 


97 


109 


121 


132 


'144 


156 




300 


106 


119 


132 


144 


157 


170 


9&16xU 


225 


93 


104 


116 


126 


138 


149 




250 


103 


116 


128 


139 


153 


166 




275 


113 


127 


141 


154 


168 


182 


9&18xl4 


225 


127 


142 


157 


172 


197 


203 




250 


141 


158 


175 


191 


208 


225 




275 


155 


174 


192 


210 


229 


238 


11&18X16 


200 


129 


145 


160 


175 


190 


206 




225 


145 


163 


180 


197 


214 


232 


«* 


250 


161 


181 


200 


219 


238 


258 


12&20xl6 


200 


157 


175 


194 


212 


231 


250 




225 


176 


197 


218 


238 


260 


281 


" 


250 


195 


219 


242 


265 


289 


312 


12&20xl8 


175 


154 


173 


191 


209 


228 


246 




200 


176 


198 


218 


238 


260 


282 


" 


225 


198 


222 


246 


268 


293 


318 


14&23X18 


175 


206 


231 


256 


280 


305 


330 




200 


236 


265 


293 


320 


349 


377 


• ' 


225 


266 


298 


329 


360 


392 


424 


15&25x20 


160 


246 


276 


306 


334 


364 


394 




180 


277 


311 


344 


376 


410 


444 


>' 


200 


308 


346 


383 


418 


455 


493 


a7&28x20^ 


160 


310 


348 


385 


420 


458 


495 




180 


349 


391 


433 


473 


515 


558 




200 


387 


434 


481 


525 


572 


619 


17&28x22 


150 


320 


359 


397 


433 


472 


510 




165 


352 


394 


437 


477 


520 


563 


" 


180 


383 


429 


476 


520 


567 


613 


19&32x22 


150 


408 


459 


508 


554 


604 


654 




165 


449 


504 


558 


612 


665 


720 


" 


180 


490 


550 


609 


675 


725 


784 



STEAM ENGINES 



319 



HIGH SPEED COMPOUND NON-CON- 
DENSING ENGINES 

Cut off at % Stroke 



Size 


T3 




Initial 


pressure 




















100 


110 


120 


13a 


140 


150 


7&11X10 . 


300 


53 


59 


65 


72 


78 


84 




325 


57 


64 


71 


79 


84 


91 


' ' 


350 


62 


69 


76 


83 


91 


98 


8«S:13xlO 


300 


72 


80 


89 


97 


106 


114 




325 


78 


87 


96 


105 


114 


124 


" 


350 


83 


94 


104 


113 


123 


133 


8&;L3x12 


250 


72 


80 


89 


97 


106 


114 




275 


79 


88 


98 


107 


116 


126 


•' 


300 


86 


96 


107 


116 


127 


137 


9&16xl2 


250 


100 


113 


125 


137 


149 


161 




275 


111 


124 


138 


150 


164 


177 


•' 


300 


121 


136 


150 


164 


178 


183 


9&16xl4 


225 


105 


118 


131 


143 


156 


169 


«* 


250 


117 


132 


146 


159 


173 


187 




275 


129 


145 


160 


175 


190 


206 


9&18xl4 


225 


144 


161 


178 


194 


210 


229 




250 


.160 


179 


198 


216 


235 


254 


" 


275 


175 


197 


218 


237 


258 


279 


ll&18xl6' 


200 


146 


164 


181 


198 


215 


232 




225 


164 


184 


204 


222 


241 


261 




250 


182 


204 


227 


247 


269 


291 


12&20xl6 


200 


177 


198 


220 


240 


260 


282 




225 


199 


223 


247 


270 


293 


318 




250 


221 


248 


275 


300 


326 


353 


1 12&20xl8 


175 


174 


196 


216 


236 


257 


278 


'■ 


200 


199 


224 


248 


270 


294 


318 


*• 


■225 


224 


251 


27 S 


304 


330 


357 


I 14&23xl8 


175 


233 


262 


290 


316 


343 


372 




200 


267 


300 


332 


362 


393 


426 


I " 


225 


300 


336 


373 


407 


442 


479 


! 15&25x20 


160 


279 


313 


346 


378 


411 


445 


'! 


180 


314 


352 


390 


426 


462 


501 




200 


349 


392 


434 


473 


514 


556 


: 17&28x20 


160 


351 


393 


436 


476 


516 


559 


' 


180 


395 


442 


490 


534 


582 


629 


" 


200 


438 


492 


544 


594 


645 


699 


: 17&28x22 


150 


362 


405 


450 


491 


532 


576 




165 


398 


446 


495 


539 


587 


634 


«• 


180 


433 


487 


538 


587 


638 


691 


19&32x22 


150 


461 


520 


576 


628 


682 


738 


i *' 


165 


509 


571 


632 


690 


750 


813 


i . " 


180 


555 


623 


690 


753 


818 


886 





320 STEAM ENGINEERING 



LOW SPEED COMPOUND CONDENSING 
ENGINES 



[ea| 



As a general rule it is not a good id 
to expand steam until the terminal pres- 
sure is lower than the back pressure, and 
while this rule limits the total ratio of 
expansion in a non-condensing compoimd 
engine to a point that will make the ter- 
minal pressure about 15 pounds absolute 
under ordinary conditions, and other 
considerations may raise it several 
poimds, the addition of a condenser to 
the plans and specifications for an en- 
gine to develop a given power, makes 
it possible to expand the steam much 
lower, as the terminal pressure may be 
as low as 6 pounds where the back pres- 
sure is about 3 pounds, both absolute or 
above a vacuum. This is accomplished 
by using a larger low pressure cylinder 
and cutting off steam earlier in the high 
pressure. 

When designing compound engines 
the same rules can be used for both con- 
densing and non-condensing service, but 
inasmuch as there is more than one 
method for this work, a better illustra- 
tion of the whole process (or a larger 
part of it), will result from using a 



STEAM ENGINES 321 

somewhat different plan here than 
was adopted for the non-condensing 

& engine. 

In order to make a complete and logi- 
cal comparison later on, it is assumed 
that 300 horse power is wanted, and that 
the stroke of the engine is 42 inches with 
a piston speed of 630 feet per minute. 
When the other data is added the whole 
appears as follows,- giving an intelligent 

j basis for designing the cylinders. 

Gauge pressure at the 

boiler = 140 pounds. 

Atmospheric pressure . . . = 15 " 

Initial pressure absolute = 150 " 

Terminal " " = 7.5 

'' Total ratio of expansion = 20 " 

Ratio of cylinders = 1 to 4 

Stroke = 42 inches. 

■ Speed =90 revolutions. 

Piston speed per minute = 630 feet. 

To develop 300 horse power. 

The first point is to determine the 
ratio of expansion for the high pressure 
cylinder. According to a rule previously 
given, this is found by dividing the total 
ratio of expansion by the ratio of cylin- 
ders. In this case it is 20-5-4=5. The 



322 STEAM ENGINEERING 

ratio of expansion for the high pressure 
cylinder is 5, therefore steam is cut ofif in 
this cylinder at yi or .20 of the stroke, 
as clearance is ignored in these estimates. 
There is a drop in pressure between the 
two cylinders, especially in the case of a 
cross compound engine due to free ex- 
pansion in the receiver, but inasmuch as 
this would seldom be the same in two or 
more engines in practice, no attempt is 
made to account for it when estimating 
cylinders for a compound engine. 

The terminal pressure in the high pres- 
sure cylinder is 150-7-5=30 pounds. 
This is also the back pressure in the 
same cylinder, and the initial for the low 
pressure, all absolute. The mean effec- 
tive pressure for the high pressure cylin- 
der is therefore — 



(1.6094+1) X 150 

■^^ —^ 30 = 48.3 pounds. 



One-half of the total of 300 horse 
power, or 150, is to be developed in this 
cylinder, therefore the required area is — 

150X33,000 ,^„^ 
— — - — — — - =162.6 square mches, 
48.3X630 ^ 

which is a circle 14.4 inches in diameter, 



STEAM ENGINES 323 

but would be called U inches in order to 
avoid impracticable fractions. The 
area is 153.93 square inches, as this cyl- 
inder is 14X42 inches. 

The given ratio of cylinders is 1 to 4, 
therefore, the area of the low pressure 
cylinder is 153.93x4 = 615.7 square 
inches, corresponding to a circle 28 
inches in diameter. 

As the initial pressure in this cylinder 
is 30 pounds and the ratio of expansion 
is 4 with a back pressure of 3 pounds, 
which is in accordance with ordinary 
practice in condensing engines, the mean 
effective pressure is — 

03863+1) X30 ^ ^ ^^ gg p^^^^3^ 

4 
Under these conditions the high pres- 
sure cylinder will develop— • 

153.93 X48.3X630 _^,,^ ^^^^^ p^^^,. 
33,000 

The low pressure will develop— 
615.7X14.89X630 _^^, ^^^^^ p^^^,. 
33,000 

The complete engine will develop 
141.9+175 = 316.9 horse power, which 



324 STEAM ENGINEERING 

is about 5 per cent, more than is wanted 
but it is not always practicable (although 
it may be possible), to design cylinders 
under given conditions so that exactly 
the desired amount of power will be 
developed, but the automatic cut oflF 
device on the high pressure cylinder 
will respond to the demand for power 
and give desired results. 

The power developed in these cylin- 
ders is not equal, but these conditions 
show the necessity of providing an auto- 
matic cut off for the low pressure cylin- 
der, which can be adjusted by the engi- 
neer, then by lengthening the comparative 
range of the governor he can lengthen 
the point of cut ofif in the low pressure 
cylinder, thus throwing some of the load 
from the low to the high pressure cylin- 
der, and make them nearer equal in this 
respect, for in these estimates it is as- 
sumed that the cut off in the low pres- 
sure cylinder corresponds to the ratio of 
areas between the two cylinders. In 
this case the ratio is 4, therefore the cut 
off in the low pressure cylinder is as* 
sumed to be at 34 or .25 of the stroke. 



A COMPARISON OF RESULTS 

Here are three engines designed to 
develop approximately 300 horse power 



STEAM ENGINES 326 

all having the same stroke, and as they 
revolve 90 times per minute the piston 
speed is alike, therefore, it is practicable 
to illustrate their relative economy by 
determining the weight of steam required 
per horse power hour by each, and com- 
paring the results. The weight of steam 
used is found by the following rule. 

Multiply the area of piston in square 
inches, by the distance in inches travelled 
by the piston at the point of cut off, and 
by the number of strokes per hour, and 
divide by 1,728. Multiply the quotient 
by the weight of steam per cubic foot at 
given pressure. Divide by the horse 
power developed, and the quotient is the 
weight of steam used per horse power 
hour. If it is desired to take clearance 
into account, multiply the above result 
by the given per cent, of clearance and 
add it to the amount. 

The following data applies to the first 
engine to be considered: 

Size of cylinder = 20 X 42 inches. 

Area " " =314.16 square inches. 

Ratio of expan- 
sion =4 

Distance to point 

of cut off = 42 -^4 = 10.5 inches. 



326 STEAM ENGINEERING 

Strokes per hour = 90X2X60 = 10,800 
Absolute pressure = 120 pounds 

Weight per cubic 

foot = .2724 pounds. 

Horse power de- 
veloped =299.88 

Then 314.16X10.5X10,800 ^ 1,728 = 
20,616 cubic feet per hour, the weight of 
which is .2724 pound per cubic foot, 
therefore the total weight used per hour 
is 20,616 X. 2724 = 5,615.798 pounds. 
Then 5,615.798 -^ 299.88 = 18.72 pounds 
per horse power hour. 

Assuming that the clearance is 3 per 
cent, of the whole volume of the cylinder 
will raise this to 18.72+ (18.72 X. 03) = 
19.28 pounds per horse power hour. 
Adding 30 per cent, to this increases it 
to 25.06 pounds, and the reason for this 
addition is found in the following lines. 

The weight of steam accounted for by 
the above calculation which is based on 
theoretical conditions, does not repre- 
sent the weight of water pumped into 
the boilers, and furthermore it is not in- 
tended for this purpose, as the only 
reliable process is to either measure or 
weigh the water as it goes into the boilers 
after which some of it will be used to run 



STEAM ENGINES 327 

the engine forming a part of the plant, 
a portion will be used for operating 
pumps and other machines, and the 
remainder will be lost through leaky 
joints, or disappear as the result of radi- 
ation. These factors are seldom or 
never the same in two different plants, 
therefore it is not practicable to make a 
fixed allowance that will apply to all 
cases, and no attempt is made to do so 
here. However, if 30 per cent, is added 
to the results secured by these calcula- 
tions, the sum will represent what can 
be realized under first-class conditions, 
but whether it is secured or not, can only 
be determined by direct experiment in 
each test, as it is made from time to 
time. As 30 per cent, is added in all 
cases mentioned here for comparison, 
the results bear the same relation to each 
other that existed previously, but the 
final amount in each case is reasonable, 
hence is an improvement over the rate 
secured by the necessarily crude and 
unfinished calculation previously se- 
cured. 

The following data applies to the sec- 
ond engine to be considered, which is of 
the compound non-condensing type: 



328 STEAM ENGINEERING 

Size of cylinders. =14 and 22X42 inches. 

Area of high pres- 
sure cyHnder. . . =153.93 square inches. 

Ratio of expansion 
for this cylinder = 3 

Distance to point 
of cut off =42 -^3 = 14 inches. 

Strokes per hour. =90X2X60 = 10,800 

Absolute pressure = 140 pounds. 

Weight per cubic 
foot = .3147 pound. 

Horse power devel- 
oped =301.2 

Then 153.93 X 14 X 10,800 4- 1,728 = 
13,468 cubic feet per hour, the weight of 
which is .3147 pound per cubic foot, 
therefore the total weight used per hour 
is 13,468 X. 3147 = 4,238.379 pounds. 

Then 4,238.379-7-301.2 = 14.07 pounds 
per horse power hour. Adding 3 per 
cent, for clearance raises this to 14.49 
pounds and adding 30 per cent, as before 
gives a total of 18.83 pounds per horse 
power hour. 

The following data applies to the 
third engine to be considered which is of 
the compound condensing type: 



STEAM ENGINES 329 

Size of cylinders = 14 and 28 X42 inches. 

Area of high pres- 
sure cylinder = 153.93 square inches. 

Ratio of expansion 
for this cylinder = 5 

Distance to point 

of cut off = 42 -T- 5 = 8.4 inches. 

Strokes per hour = 90 X 2 X 60 = 10,800 
•Absolute pressure = 150 pounds. 

Weight per cubic 
foot = .3358 pound. 

Horse power de- 
veloped =316.9. 

Then 153.93 X 8.4 X 10,800 -r- 1,728 = 
8,081 cubic feet per hour, the weight of 
which is .3358 pound per cubic foot, 
therefore the total weight used per hour 
is 8,081 X .3358=2,713.6 pounds. 

Then 2713.6 4-316.9 = 8.55 pounds per 
horse power hour. Adding 3 per cent, for 
clearance raises this to 8.8 pounds, and 
adding 30 per cent, as before gives a total 
of 11.44 pounds per horse power hour. 

The difference between results that 
can be secured by use of these engines 
does not seem large as here stated ac- 
cording to common practice, but when 
it is multiplied by the power developed 
in a given case, and this goes on from 10 
to 24 hours per day according to the 
service in which an engine is used, the 
economy of securing a first-class engine 



330 STEAM ENGINEERING 

and of placing a competent engineer in 
charge of it, will be very plain. 

The next four tables give sizes of com- 
pound condensing engines that will ren- 
der satisfactory service under given 
conditions. The first contains engines 
suitable for 125 poimds boiler pressure, 
when the load causes about 14 expansions 
and consideration of this matter calls for 
explanation of what constitutes unrea- 
sonable conditions. The fact that a com- 
pound engine must run under conditions 
which give a low terminal pressure if 
economical results are desired, has been 
made plain on previous pages, but the 
prospective customer, or the superin- 
tendent or the owner of a plant where 
one or more of these engines are to be 
used, frequently forget that a low ter- 
minal pressure means a comparatively 
light load. Now, if a very heavy load 
is put on a compound engine, resulting 
in a high terminal pressure in the second 
cylinder, this improved kind of an en- 
gine will be almost as wasteful as a sim- 
ple engine, when used under similar 
conditions. 

When the manufacturer of a com- 
pound engine is asked to state the 
amount of power that it will develop in 
an emergency, he usually answers the 
question just as it is stated, although he 
knows that while carrying the maximum 



STEAM ENGINES. 331 

load it is wasteful of steam. If a steam 
user proceed? to put this load on his 
engine for 10 hours or more per day, and 
the result is very unsatisfactory, as the 
wear and tear is excessive, and the coal 
bill proves to be very large, he ought to 
blame nobody but himself. He should 
ask the engine builder to state the power 
that it will develop on an economical 
basis, and then keep within the limits so 
determined. 

The second table contains sizes of low 
speed engines where the ratio of cylin- 
ders is 1 to 4. This means that the area 
of the low pressure is four times as much 
as the high pressure. The ratio of diam- 
eter is 1 to 2 in all such cases, as stated 
in the list given. These engines are suit- 
able for 150 pounds pressure by the gauge. 

The third table contains the proper 
sizes for given loads where the pressure 
is 120 pounds, and the speed is high. 
The ratio of cylinders is about 1 to 3.5 
although it varies with the larger sizes 
given. 

The fourth table contains sizes of 
cylinders with a ratio of 1 to 4, to carry 
150 pounds pressure. In this and in the 
^preceding table three different speeds 
are given with power developed accord- 
ingly. If the speed of an engine is in- 
creased it will develop more power with 
out changing it from an economical to a 



332 



STEAM ENGINEERING 



wasteful machine, provided that the 
boiler pressure and the point of cut off 
remain constant. Excessive speed will 
cause rapid wear, and better lubrication 
will be necessary, but these details can 
Usually be dealt with properly if an in- 
telligent engineer is in charge and his . 
recommendations are adopted. The 
speeds given in these tables are not ex- 
cessive, and there is no good reason why 
a well made engine should give trouble 
of any kind while developing power as 
stated in them. Good judgment should 
be shown in the selection of a superior 
cylinder oil for all engines working under 
high pressure. The cost per galloii" 
should be a secondary consideration, 
J)rovided it does satisfactory work. 

COMPOUND CONDENSING ENGINES. 

Long stroke. Low speed. About 14 Expansions . 

Ratio of Cylinders Approximately 1 to 3Vi. 



Fo] 


■ 125 pounds boiler pressure. 




Size 


'X3 


t 


Size 


73 
(U 


1 




a 


o 




a 






03 


pL, 




w 


Oi 


9&16x30 


100 


96, 


22&40x54 


70 


750 


10& 18x30 


100 


122 


23&42x60 


65 


864 


ll&20x36 


90 


163 


24&44x60 


65 


948 


12&22X36 


90 


190 


25&46x60 


65 


1,036 


13&24x36 


90 


234 


26&48x66 


55 


1,050 


14&26x42 


80 


285 


28&50x66 


55 


1,140 


16&28x42 


80 


330 


29&52x66 


55 


1,233 


I7&30x42 


80 


380 


30&54x72 


50 


1,319 


18&32X48 


75 


463 


31&56x72 


50 


1,411 


19&34x48 


75 


523 


32&58x72 


50 


1,620 


20&36x54 


70 


615 


34&62x72 


50 


1.739 


21&38x54 


70 


678 


36&64x72 


50 


1,858 



STEAM ENGINES 



333 



COMPOUND CONDENSING ENGINES. 



Long stroke. Low speed. 16 Expansions. 
Ratio of cylinders 1 to 4. 



For 150 pounds boiler pressure. 




Size 


♦a 


1 


Size 


a 


1 




w 


a, 




w 


0^ 


8&16x30 


inn 


99 


20&40x54 


70 


784 


9&18x30 


100 


12G 


21&42x60 


65 


891 


.0&20x36 


90 


168 


22&44x60 


65 


978 


:.l&22x36 


90 


203 


23&46x60 


65 


1,069 


.2&24x36 


90 


242 


24&48x66 


55 


1.084 


13&26X42 


80 


294 


25&50x66 


55 


1,176 


14&28x42 


80 


341 


26&52x66 


55 


1,272 


15&30x42 


80 


392 


27&54x72 


50 


1,360 


16&32X48 


7.5 


478 


28&56x72 


50 


1,463 


17&34x48 


75 


539 


29&58x72 


50 


1,569 


18&36x54 


70 


634 


31&62x72 


50 


1,793 


19&38x54 


70 


707 


32&64x72. 


50 


1,900 



COMPOUND CONDENSING ENGINES. 

Short stroke. High speed. About 14 Expansions. . 
Ratio of Cylinders approximately 1 to 3.5. 



For 120 pound 


s boiler pressure. 






•d 


u 


•a 


V 


TS 


S3 


Size 


^ 


^ 




^ 


n^ 


% 














o 




w 


fu 


w 


&i 


CO 


Oh 


7&llxl0 


300 


45 


325 


49 


350 


53 


8&13xl0 


300 


61 


325 


66 


350 


71 


8&13xl2 


250 


61 


275 


67 


300 


73 


9&16xl2 


250 


84 


275 


93 


300 


101 


9&16xl4 


225 


89 


250 


98 


275 


108 


U&18xl4 


225 


122 


250 


136 


275 


149 


ll&18xl6 


200 


124 


225 


1.39 


250 


155 


12&20X16 


200 


150 


225 


169 


250 


187 


12&20X18 


175 


148 


200 


169 


225 


190 


14&23X18 


175 


198 


200 


227 


225 


255 


15&25x20 


160 


236 


180 


266 


200 


295 


17&28x20 


160 


298 


180 


335 


200 


372 


17&28x22 


150 


308 


165 


337 


180 


368 


19&32x22 


150 


391 


165 


432 


180 


470 



334 



STEAM ENGINEERING 



COMPOUND CONDENSING ENGINES 

Short stroke. High speed. 16 Expansions 
Ratio of cylinders 1 to 4. 

For 150 pounds boiler pressure 





•a 


u 


-0 


u 


13 


u 


Size 


1 


1 


I 


o 


1 


1 




w 


cu 


w 


CLi 


W 


D^ 


SJ^&llxlO 


300 


42 


325 


45 


350 


49 


6}^&13xl0 


300 


58 


325 


63 


350 


68 


6i^&13xl2 


250 


58 


275 


64 


300 


70 


8 &16xl2 


250 


88 


275 


97 


300 


106 


8 &16xl4 


225 


93 


250 


103 


275 


113 


9 &18xl4 


225 


117 


250 


130 


275 


143 


9 &l8xl6 


200 


119 


225 


134 


250 


149 


10 &20xl6 


200 


147 


225 


165 


250 


184 


10 &20xl8 


175 


145 


200 


166 


225 


186 


12 &24xl8 


175 


199 


200 


228 


225 


256 


12J^&25x20 


160 


240 


180 


270 


200 


300 


14 &28x20 


160 


288 


180 


324 


200 


360 


14 &28x22 


150 


297 


165 


326 


180 


356 


16 &32x22 


150 


387 


165 


427 


180 


465 



PROVIDING FOR INCREASE OF POWER 

WHEN INSTALLING A COMPOUND 

ENGINE 

When a steam plant is to be designed 
constructed and installed, .the serviced 
of a competent designing engineer should 
be secured to draw the specifications 
and superintend the work, as it is im- 
possible to lay down rules for such cases 
that will cover every contingency that 
may arise without making them too 
cumbersome for practical use. It is pos- 
sible and practicable, however, to make 
a few suggestions and give rules that 



STEAM ENGINES 335 

apply to this work, that will prove valu- 
able to both engineer in charge, and 
owner of the plant, as they either give 
clear directions, or point out the neces- 
sity for investigation along these lines. 

Suppose that a cross compound con- 
densing engine is preferred, and for one 
or two years the load is to be very light 
after which it will be greatly increased. 
This engine is illustrated in Fig. 42, but 
while the load is light the low pressure 
connecting rod may be disconnected 
from the crank pin, and the eccentric 
straps taken off, thus leaving it a sim- 
ple non-condensing engine. In order to 
illustrate the matter, assume that a 
single diagram taken from this engine 
under these conditions appears like Fig. 
43. It is now working under economical 
conditions as the terminal pressure is not 
enough above the atmosphere to denote 
serious waste, yet it is sufficient to pre- 
vent excessive cylinder condensation. 
This engine was designed for 150 pounds 
boiler pressure by the gauge, but only 100 
is carried now, as no more is needed, and 
it is useless to carry a high boiler pres- 
sure unless it can actually be used to 
good advantage. 

Fig. 44 is another diagram from the 
same engine while carrying 100 pounds 
boiler pressure, but the load is much 
heavier than before. This is a wasteful 



336 STEAM ENGINEERING 




FiG>.4 2 



FiS.A-b 



STEAM ENGINES 



337 




r i ii'— T^ 




■=^^=w^ 


1 




\ 




F\e^A^^ 



338 STEAM ENGINEERING 

condition of affairs, therefore, the boiler 
pressure is raised to 150 pounds and the 
diagram Fig, 45 is taken. The in- 



creased boiler pressure has shortened 
the point of cut off and lowered the ter- 
minal pressure until economical condi- 
tions are again secured. However, the 
load is again increased until it produces 
a diagram similar to 46. The limit of 
boiler pressure is shown by the diagram, 
therefore, the next move is to connect 
the low pressure connecting rod, adjust 
the eccentric straps, make proper steam 
connections, put the condenser into ser- 
vice and run the engine as it was origi- 
nally intended to be used. A diagram 



STEAM ENGINES 



339 



from the high pressure cyHnder as now 
running, is shown in Fig, 47, and one 



from|the low pressure cylinder appears 
in Fig. 48. 



340 



STEAM ENGINEERING 



The flexibility of an engine of this j 
kind is well illustrated by this example, 
and there are other points that are ; 




FIG. ^8 



worthy of special attention. The com- 
plete engine should be installed at first, 
and such parts as are not wanted for j 
immediate use can be held in reserve 
until needed. While this is the best 
plan that can be devised, it is objected 
to by some engineers and owners, as j 
they believe that only the high pressure 
side should be installed at first, as the 
low pressure side can be added when 
wanted, thus saving interest on the in- 
vestment, etc. This plan is earnestly 
objected to for. the following reason: 
My observation and experience teaches 
that although a plant may be started in 
an unfinished condition, it will be run 
that way as long as possible. 

The engine may be overloaded until 
the speed is much less than it ought to 



STEAM ENGINES 341 

be, and the terminal pressure is so high 
that a large quantity of heat is thrown 
away every day, but so long as the plant 
continues to run without an absolute 
stop, there is no time to shut down for 
improvements, money cannot be spared 
for that purpose, etc., and matters usual- 
ly remain in that condition until enough 
money has been wasted to more than 
pay the cost of needed improvements, 
after which they are made, resulting in 
marked economy of operation causing 
the owner to wonder why he did not do 
it before. 

Other reasons why the low pressure 
side of this engine ought to be installed 
complete when the plant is first erected, 
are that if it is not done then, the firm 
who build these engines may go out of 
business, their works may burn down, 
or they may be so busy when the extra 
parts are wanted that it will be necessary 
to wait a year for the work to be com- 
pleted. 

■ On the other hand if the low pressure 
side is completed, it may be put into 
service at a few hours' notice, thus saving 
serious loss due to shutting down the 
plant for several weeks, for such ma- 
chinery is sometimes wanted when least 



342 



STEAM ENGINEERING 



expected. Suppose that the piston rod 
on the high pressure side should break, 
allowing the piston to be forced out 
through the cylinder head, taking pieces 





F«e|. 



of the cylinder with it, thus making it 
necessary to build a new cylinder, piston 
and piston rod. The high pressure con- 
necting rod could then be disconnected, 
and the eccentric straps taken off. The 



STEAM ENGINES 



343 



piping might be quickly changed to 
deliver steam directly to the low pres- 
sure cylinder, the boiler pressure reduced 
to 30 or 40 pounds and in a few hours 




the plant could be in full operation, 
which might easily be continued indefi- 
nitely, or until repairs were completed. 
This would also make it unnecessary to 
work overtime on the job, thus saving 



344 STEAM ENGINEERING 

heavy expense, as overtime must be 
paid for extra, although the work cannot 
be carried on to good advantage at night 
with ordinary lighting facilities, by men 
who are taxed beyond their strength in 
working from 24 to 48 hours without 
rest. 

Another advantage of this plan is that 
it does not include expensive changes in 
the engine in order to secure the full 
power, no additions are required, and 
the speed is not increased, therefore it is 
not necessary to provide a new and 
larger main pulley on the jack shaft to 
keep the machinery speed ' constant 
while the engine speed is changed, and 
the main belt does not have to be length- 
ened to meet new conditions. 

Fig. 49 illustrates a double tandem 
• compoimd condensing engine that is a 
very fiejyble unit. If only a small por- 
tion of the full load is to be carried at 
first, one side can be laid up, and a light 
boiler pressure will be sufficient for the 
other. After six months have elapsed 
the other side can be started and each 
run alternately, thus keeping one side 
in reserve, ready for use in case of acci- 
dent to the other. As the load is in- 
creased, the boiler pressure can be 




;r frame type 




CROSS COMPOUND HARRIS-CORLISS ENGINE, STANDARD GIRDER FRAME TYPE 



STEAM ENGINES 345 

raised, and when it is no longer econom- 
ical to run with one side they can both 
be used with a low boiler pressure, then 
as more load is added it can be raised 
imtil the full power of the complete 
engine is utilized. 

While the highest efficiency cannot be 
expected imder all of these varying con- 
ditions, it will be realized imder some 
of them, hence the total difference is 
small when compared with the benefits 
secured. 



SECTION 4 

PRACTICAL POINTERS FOR PRO- 
GRESSIVE POWER PLANT 
OPERATORS 

PACKING A MAN-HOLE COVER 

Putting in a man- hole cover is a rough 
job, yet it requires a certain amoimt of 
care in order to make it successful, for 
if the gasket is cut in tv/o by coming into 
contact with a sharp corner it will cause 
a bad leak, and where a tubular gasket 
is used, it is necessary to clamp both 
ends of it in proper place, or else it will 
be necessary to empty the boiler and 
make a new joint. 

While this is very unpleasant to all 
concerned, when there is plenty of time to 
repack the joint, it is much worse to dis- 
cover such a mistake about an hour be- 
fore it is time to start the engine, because 
every self-respecting engineer wishes to 
start his plant on time, as failure to do 
so, even when there is apparently a good 
reason for it, injures his reputation and 
causes his employer to lose confidence in 
him. On this account it is always a 
good idea to raise pressure on a boiler 
after it has been cleaned and inspected, 
before it is wanted for actual service, then 
if it becomes necessary to cool it off and 



PRACTICAL POINTERS 347 

repack one or more joints it will cause 
but little expense and no delay to the 
plant. 

FAILURE OF MAN-HOLE GASKETS 

Sometimes a man-hole or a hand-hole 
gasket will blow out under pressure, but 
such accidents are always due to bad 
design of the parts, carelessness in mak- 
ing the joints, or failure to properly care 
for them. The surfaces to be packed 
may be roughened by corrosion, or made 
winding by abuse of the boiler. It may 
be necessary in such a case to use two 
flat gaskets, but tjiis should be avoided 
as long as possible, because the extra 
thickness of gasket makes it more liable 
to blow out. When a boiler is filled 
with cold water and one of these joints 
leaks, it is a great temptation to seize a 
long wrench and screw the nut as far 
down as possible. Of course this is 
proper to a limited extent, but if exces- 
sive leverage is applied the cast iron dog 
which spans the hole may be broken. 
It is well to have an extra one on hand 
for such emergencies. 

PACKING A HAND-HOLE COVER 

Horizontal return tubular boilers are 
frequently made without a hand hole in 
the rear head. As long as the boiler is 



348 STEAM ENGINEERING 

in service, there is no need of a hand hole 
here, but when it is laid off to be cleaned 
it is very convenient to have a place 
where at least a little fresh air can be 
secured, also an opening through which 
tools can be passed. It is difficult in 
some cases and impossible in others to 
properly care for a hand-hole cover ow- 
ing to its location. When packed with 
a rubber gasket and the nut is screwed 
down firmly, the joint may be tight, but 
when steam is raised the heat softens the 
gasket, and pressure acting on the back 
of the cover settles it down into place 
until the nut is quite loose. If it is 
made tight again before all pressure is 
taken off, it will probably make a good 
joint until it is time to remove the cover 
again, but if there is no chance to do this 
for several days or weeks, and the steam 
pressure goes down to zero, it will prob- 
ably leak, and water coming through 
the imperfect joint will cause corrosion 
that in course of time will waste away 
the head until a tight joint cannot be 
secured. 

FITTING A HAND-HOLE COVER 

Before attempting to pack a hand- 
hole cover for the first time, on either an 
old or a new boiler, the engineer should 
put the cover in place without packing 



PRACTICAL POINTERS 349 

and be sure that the hole is large enough 
to allow the cover to rest firmly on the 
head without binding on the edges as 
otherwise the gasket may not be clamped 
firmly, and when subjected to pressure 
it may be blown out. After a tight 
joint is secured, and all lost motion due 
to shrinking of the gasket has been taken 
up, if the bolt projects through the nut, 
another should be screwed on to protect 
the end of it, then the dog, nuts and bolt 
should be covered with asbestos to pro- 
tect them from the fierce heat. 

RUBBER AND METAL GASKETS 

As a general rule, to which there may 
be a few exceptions, a rubber man-hole 
gasket can be used several times, provid- 
ed there is a broad surface in contact to 
afford a good bearing. If it is desired to 
use a gasket more than once, the side put 
next to the cover should not contain 
anything to prevent it from sticking to 
the metal, but the other side ought to be 
coated with Dixon's graphite, ground 
■fine and mixed with cylinder oil to the 
consistency of a thick paste. This will 
prevent the rubber from sticking to the 
head, and if care is taken to clean pieces 
of scale, etc., off from the surfaces to be 
packed, and the cover is put back in 
exactly the place from which it was re- 



350 STEAM ENGINEERING 

moved, it should make a tight joint. '■] 
This cannot be done with some of the 
modern boilers where the head is simply- 
flanged inward and planed off to receive 
the cover, as the surface is too narrow 
to admit of it. The metal cuts into the 
gasket until it is practically metal to 
metal, with only enough rubber between 
them to fill the places that would other- 
wise be vacant, therefore, such a joint 
will not readily fail under pressure. 
Lead gaskets are sometimes adopted for 
this service, and they would be used 
much more if it was possible to put them 
back after cleaning and inspecting boil- 
ers, and secure tight joints with little 
trouble, but as a rule this is not practi- 
cable owing to the difficulty of getting 
them back exactly in their former posi- 
tions. 

LOCATING FUSIBLE PLUGS 

The fusible plug in a horizontal tubu- 
lar boiler should be located in the rear 
head, three inches above the upper tubes. 
As such a boiler ought to be one inch 
lower at the rear than at the front, in 
order to have all water drain to the blow- 
off pipe, this will leave but two inches of 
water above the highest part of the tubes 
when the fusible plug is uncovered, and 
this is no more than enough to insure 
safety. 



PRACTICAL POINTERS 351 

In a locomotive boiler it should be in 
the highest part of the crown sheet, and 
in every other kind it ought to be located 
where it will melt and give emphatic 
notice of the condition of affairs, before 
any part of the boiler is injured through 
lack of water. 

LEVELLING BOILERS 

The upper row of tubes in a tubular 
boiler, should be set level crosswise of 
the boiler, causing the water to cover 
them evenly, then if the dome or the 
steam nozzles do not stand plumb, the 
defect may be counteracted by using 
pipe flanges that are thicker on one edge 
than the other. 

HEAVILY LOADED BOILERS 

Boilers are frequently made to devel- 
op more power in practice than their 
builders intended, and there are engi- 
neers who claim that it is a good idea to 
secure as much power as possible from a 
boiler, and then put it in the junk pile. 
One reason why this plan should not be 
followed is that such boilers will not be 
put into the junk pile soon enough, but 
will be used long after they become 
unsafe, as it is an expensive job to shut 
down a plant, remove old boilers and 
install new ones. 



352 STEAM ENGINEERING 

SMOKE CONSUMERS AND PREVENTERS 

A good smoke consuming device is a 
valuable apparatus, but a smoke pre- 
venter is much better. A very good 
device for this purpose is a large furnace, 
and plenty of boiler power, enabling the 
fireman to run comparatively slow fires, 
and thus consume the coal properly, 
leaving no valuable carbon to go up the 
chimney. It naturally follows that a 
good fireman will be necessary in order 
to secure the full advantage of furnaces 
and boilers that are well proportioned 
for the service that they were intended 
to perform. 

A boiler should be high enough above 
the grates to permit proper combustion 
of the coal, before the flames come into 
direct contact with the metal, because 
this contact lowers the temperature, 
which must be very high in order to se- 
cure good results. Under these condi- 
tions a fire may be carried 8 or 10 inches 
thick, and if coal is supplied frequently 
and in small quantities, with a suitable 
amount of air admitted above the fire, it 
is possible to run a plant without having 
much black smoke come out of the 
chimney. 

A certain plant with which the writer 
is somewhat familiar, sends large quan- 
tities of very black smoke out of its 



PRACTICAL POINTERS 353 

chimneys. One reason for this is that 
although the boilers are rated fully as 
high as they ought to be, still they are 
caused to develop 50 per cent, more 
power than they were built for. This 
leaves the fireman no chance to manage 
his fires to prevent smoke by firing in 
front and allowing the smoke producing 
elements to be consumed by passing 
over the rest of the fire, as he must man- 
age his fires to produce the greatest pos- 
sible amount of steam without regard to 
good appearance or economy of fuel. 

FORCED DRAFT 

When forced draft is mentioned it 
may mean that boilers under which it is 
used are worked beyond their rated 
capacity, but it may be used to force air 
through a bed of very small coal, that 
could not be used with ordinary natural 
draft. In the latter case it is not wise 
to claim that the forced draft is ruining 
the boilers of a plant, because no more 
air passes through such a furnace, and 
the temperature is no higher in it than 
would be secured by burning large coal 
with good natural draft. 

TEMPERATURE OF GASES 

Whether a boiler is overloaded or not 
can usually be determined by observing 



354 STEAM ENGINEERING 

i 

the temperature of the products of com- ; 
bustion as they pass to the stack. If , 
they are less than 100 degrees higher 
than the temperature of steam at the 
pressure carried, the boiler cannot be 
seriously overloaded, but if they are 200 
or 300 degrees higher, it indicates that 
more power is demanded than the boiler 
can supply at an economical rate. This 
conclusion is based on the fact that if 
the grate and the heating surface are 
properly proportioned, all heat generated 
on the grate will be absorbed by the 
heating surface, except enough to cause 
good draft. This applies to ordinary 
power plants, and if an exception can be 
found to this general rule it denotes 
special conditions. 

THROTTLING VS. AUTOMATIC ENGINES 

Suppose that a certain engine, fitted 
with a D slide valve and a throttling 
governor, is running with no cut off, or 
in other words it takes steam constantly 
throughout every stroke, and the load 
calls for a mean effective pressure of 40 
pounds. Under these conditions there 
will be 40 pounds pressure by the 
gauge at the beginning of each stroke, and 
this will be continued to the end. Sup- 
pose that the size of this engine and the 
speed at which it runs are such that it 



PRACTICAL POINTERS 355 

now develops 100 horse power. The 
boiler pressure is 80 pounds by the gauge 
and the reduction is made by the throt- 
tling governor with which it is fitted. 

The cylinder and governor are re- 
moved, and another cylinder of the same 
size is substituted, but it is fitted with an 
automatic cut off valve gear. The speed 
remains the same, and as the load is not 
changed it requires 40 pounds mean 
effective pressure to drive it. With a 
boiler pressure of 80 pounds arid the cut 
off taking place at yi stroke, the desired 
result is secured provided all conditions 
are favorable. 

In the first case the cylinder was filled 
with steam flowing directly from the 
boiler through the full stroke. In the 
second case the flow of steam is cut off at 
]/i stroke and no more enters during that 
stroke. Do we get the same power by 
taking from the boiler one-quarter of the 
amount taken before, thus saving 75 per 
cent, of the coal bill? No, decidedly 
not. Provided there were no leaks in 
the former case, the change will result in 
saving about 25 per cent, of fuel. Why 
is not a greater saving realized? The 
following explanation will make this 
clear: 

When one pound of water at its great- 
est density is turned into steam, it makes 
284.5 cubic feet at 80 pounds gauge pres- 



356 STEAM ENGINEERING 

i 

sure. When this steam passes the ' 
throtthng governor its pressure in this 
case is reduced to 40 pounds. The same 
weight of steam now fills 475.9 cubic feet 
of space, therefore, it fills the cylinder 
nearly twice as many times as it would 
at 80 pounds. Furthermore, the throt- 
tling process slightly superheats the 
steam, thus making it more valuable 
than before. As the pressure remains 
the same throughout the stroke, there is 
no change of temperature to cause cylin- 
der condensation, and this condition 
prevents much loss of heat. 

The automatic engine takes enough 
steam at 80 pounds gauge pressure to fill 
the cylinder one-quarter full (which 
would fill it nearly one-half full at the 
lower pressure) at each stroke, and as 
the supply is then cut off it begins to 
expand, and as the pressure is lowered 
the temperature becomes less until the 
stroke is completed. As soon as the 
crank has passed the center at that end, 
the cylinder receives another charge of 
steam at nearly boiler pressure, but 
more or less of it is condensed by raising 
the temperature of the cylinder and this 
is a direct loss that is repeated at every 
stroke. These facts are sufficient to 
explain why the automatic cut off en- 
gine is not more economical than it real- 
ly is in practice. A saving of 25 per 



PRACTICAL POINTERS 357 

cent, is sufficient to warrant its use, but 
it appears to the casual observer as if it 
should be more. 

DRAINING THE STEAM PIPE 

When it is nearly time to start an en- 
gine, do not forget to open the small 
drip valve with which every engine is, 
or ought to be fitted, for the purpose of 
allowing all water to drain out of the 
steam pipe. Everybody knows that this 
ought to be done, but when something 
happens to interfere with the regular 
routine usually observed in starting the 
plant, it may be forgotten, and if the 
throttle valve is opened when the steam 
pipe above it is full of cold water, some- 
thing interesting may happen. It may 
not amount to more than a heavy, sharp 
pound in the pipe that an engineer does 
not readily forget after he has heard it 
once, or a gasket may be blown out of a 
packed joint. It is quite possible for a 
valve to be burst in this way, causing 
the plant to be shut down until a new 
one is put in. Some of these drip pipes 
are too large, therefore if the valve in 
one of them is opened wide it will cause 
water hammer in the main pipe. A one- 
half inch drip pipe is large enough for a 



358 STEAM ENGINEERING 

12 inch main steam pipe, then there will 
not be much danger of trouble on this 
account, even if the valve is opened wide. 

WARMING THE CYLINDER. 

After all water is removed from the 
main steam pipe, the cylinder ought to 
be gradually warmed by admitting 
steam to it, taking care not to admit 
enough at first to start the engine before 
it is time to set the machinery in motion. 
If the engine is fitted with a detachable 
valve gear, this feature can be utilized 
to admit steam to alternate ends of the 
cylinder, but as soon as it is well heated, 
the valve gear should be hooked in and 
steam enough admitted to start the en- 
gine slowly. The practice of rocking 
the valve gear of an engine back and 
forth when starting it, does not tend to free 
the cylinder of water, but on the contrary 
it closes the exhaust valves before they 
would be otherwise, hence water is pre- 
vented from escaping freely. Some 
engine builders now make their valve 
gears so that they cannot be unhooked, 
as they desire to compel the engineer to 
start his engine without giving it the 
shocks and jars so frequently resulting 
from the reversing process. It is 
claimed that fly wheels are sometimes i 
loosened on the crank shaft by it. 



PRACTICAL POINTERS 359 

RUNNING AN ENGINE "OVER" 

When you write to an engine builder 
or a firm that was organized for the pur- 
pose of building engines of various sizes, 
and ask them if their horizontal engines 
will run "under" as well as "over" you 
must always expect a reply in the affirm- 
ative. There are a few engines that 
were made to run "over" and they can- 
not be reversed, but with these excep- 
tions all others are advertised to run as well 
in one direction as the other. 

It is a well known fact that when an 
engine runs "over" (or in other words, 
when the top of the fly wheel travels 
from the cylinder), the crosshead is held 
down on the guides, hence it is only nec- 
essary to adjust it to the proper height 
to bring it into line, as it is not necessary 
to make fine adjustments to prevent it 
from being lifted. 

On the other hand if an engine runs 
"imder" and is carrying a load that 
equals one-half of its rated capacity or 
more, the crosshead will be lifted at 
every stroke unless it is held firmly in 
place. 

ADJUSTING THE CROSSHEAD 

I once had charge of a Corliss engine 
that nm "under" for 132 hours per week. 
When it was comparatively cold I would 



360 STEAM ENGINEERING 

carefully adjust the crosshead between • 
the upper and lower guides until it was a ■ 
perfect fit. After it had run for an hour 
the vertical space between the guides 
near the cylinder, would be greater than 
it was before, hence the crosshead would 
. lift and cause a pound at every revolu- 
tion which was much worse on Saturday 
than it was on Monday, but the best 
engineer in the world could not stop it, 
because he could not keep the guides 
parallel vertically, when one end was hot 
and the other cool. If that fly wheel 
had revolved in the opposite direction, 
the crosshead could have been run for 
months without adjustment, and if 
there wa^ room enough between the top 
of it and the upper guide to move a piece 
of thick writing paper, there would have 
been no pound, as it would still have run 
quietly. 

SHUTTING DOWN AN ENGINE 

When an engine is to be shut down for 1 
a month or two, do not simply close the 
throttle valve and let it stand without 
further attention, until it is wanted 
again, as rust and corrosion will damage 
it much more than it would be worn 
while in use under fair conditions for a 
year, and perhaps for double that time. 
The valve or valves should be taken out 



PRACTICAL POINTERS 361 

after all packing has been removed from 
the stems, thoroughly inspected, and if 
repairs are needed, now is a very good 
time to have them made. If these parts 
are in good order, cover them with cylin- 
der oil and put them in place ready for 
use, but do not repack the stems imtil 
they are wanted for service. 

PREVENTING CORROSION 

Remove all packing from the piston 
rod stuffing box, clean all parts and 
cover them with cylinder oil, taking care 
to have a good coat of it between the 
gland and the rod. Take off the cylin- 
der head, remove the follower bolts, the 
plate, and every part of the piston. 
Carefully inspect them for defects, and 
if springs are worn, studs corroded or 
jam nuts loose, have such defects re- 
paired or the parts renewed while there 
is plenty of time for the work. Measure 
the cylinder and if it is worn much larg- 
er in one or more places than in others, 
it should be re-bored and new packing 
rings put into the piston. If this is not 
necessary, put the piston rod exactly in 
the center of the cylinder, replace every 
part of the piston ready for use, taking 
care to cover them all with cylinder oil 
at the proper time, also every square 
inch of the internal surface of the cylin- 



362 STEAM ENGINEERING 

der, in order to prevent it from rusting, 
and replace the cylinder head ready for 
use. 

STARTING AN ENGINE. 

When an engine is to be started after 
remaining idle for several months, the 
work necessary to be done depends on 
how the several parts were cared for 
when the machine was put out of com- 
mission. If the preceding directions 
concerning this important matte;- were 
intelligently followed, then it is only 
necessary to repack the valve stems and 
the piston rod, turn on the steam and 
start the engine. 

USING A STOP MOTION. 

On evjery Corliss engine, and on sev- 
eral other kinds there is or can be, some 
kind of a "stop motion," which is a 
device for preventing the engine from 
becoming a wreck in case the governor 
belt breaks or runs off from the pulleys. 
A certain class of engineers appreciate 
the value of such a device, keep it in 
good order and frequently test it to 
know that it will do the work for which 
it was installed. Another class seem to 
resent the idea that any such safeguard 
is necessary on a machine of which they 
have charge, consequently it is either 



PRACTICAL POINTERS 363 

removed altogether, or disabled and 
rendered worthless. 

This is a serious mistake, because an 
engineer n^ever knows when a hidden 
flaw will cause one of his governor pul- 
leys to collapse, neither can he tell when 
the governor belt will fail or come off, 
thus admitting the full force of steam to 
the cylinder, increasing the speed of the 
fly wheel in a few seconds until the safe 
limit is exceeded, and unless steam is 
quickly shut off a bad wreck will result. 
The cautious engineer will always take 
advantage of every possible precaution 
and keep his plant ready for a test by 
interested parties at any time. 

SPEED OF A GOVERNOR 

Do not forget that the governor on 
vour engine was built to revolve at a 
lertain speed, which is the standard for 
your case. This applies to a fly ball 
governor without springs, or adjustable 
weights. If you expect to increase the 
speed of your engine by changing the 
diameter of the governor pulley, or the 
other one, do not state that you intend 
to run your governor faster in order to 
secure the desired result, because the 
speed of it will not be changed. When 
making calculations to determine the 
diameter of a pulley required to change 



364 STEAM ENGINEERING 

the speed of an engine, just assume that 
the governor is the prime mover and that 
its speed does not change. This will 
prevent making statements that are not 
wise, and facilitate the actual work re- 
quired by the contemplated change. 

TROUBLE WITH A CRANK PIN 

If the crank pin of an engine heats 
when lubricated with good oil, it should 
be thoroughly examined for defects in 
the pin and its boxes. If none are found 
.the engine ought to be indicated without 
delay, as the sole cause of trouble may 
be located at the valves, for if they are 
not set to admit, exhaust and compress 
steam properly, the effect may appear 
in connection with the crank pin, al- 
though it is as far away as possible. 

LINING AN ENGINE 

Your engine may have been' in line 
when it was erected ten years ago, but 
that does not prove it to be so now, as it 
is quite possible for it to be changed a 
great deal, by taking up lost motion and 
making other adjustments from time to 
time, consequently if it pounds now 
more than it did when new, take out the 
piston and remove all other parts neces- 
sary, then draw a line exactly through 



PRACTICAL POINTERS 365 

the center of the cylinder and see if the 
crosshead travels parallel' to it, also if 
the crank shaft stands at right angles. 
You may be surprised at the result of 
your investigations. 

FILING CRANK PIN BRASSES 

We occasionally find an engineer who 
believes that every time his crank pin 
boxes are keyed up to take out lost mo- 
tion, they must be filed so that they will 
come together nicely at both top and 
bottom, but this is not necessary. On 
locomotives it is advisable to do this, but 
it is not required on stationary work, for 
if they are open 1/16 inch at these points 
it can do no harm, and it is much more 
convenient to put a pair of boxes on a 
planer and take off enough to answer for 
a year or two, than to file it off every 
time that lost motion must be taken up. 

AVERAGE PRESSURE OF STEAM 

The average presstue of steam acting 
on the piston of an engine is never used 
when calculating the indicated horse 
power of it, although it is so stated in ' 
some of the rules given for this purpose. 
The average pressure is always taken 
above a perfect vacuum and the back 
pressure can have no effect on it, be- 



366 STEAM ENGINEERING 

cause the two pressures are acting on 
opposite sides of the piston. The aver- 
age pressure shows the total force acting 
on the face of a piston to propel it for- 
ward, but gives no clue whatever to the 
division of the opposing force. When 
calculating this pressure it makes no 
difference whether a condenser is used 
or not, as only the initial pressure and 
point of cut off (including clearance), are 
taken into consideration. Of course, 
there are times when we want to know 
how much back pressure opposes the 
advance of an engine piston, also how 
much of it is above atmospheric pres- 
sure and other partictilars, but this only 
occurs when we wish to determine the 
mean effective pressure, and when this 
is known it is used in calculating the 
indicated horse power developed by an 
engine. 

KEEPING AN ENGINE CLEAN 

When a certain engine is correctly 
adjusted and properly lubricated, it 
will give good results, whether the cylin- 
der head is kept bright or allowed to 
"become dingy and dirty. There is, 
however, a connection between the out- 
ward appearance and the inward condi- 
tion of an engine, that is well understood 
by engineers and steam users, for as a 



PRACTICAL POINTERS 367 

general rule to which there may be a 
very few exceptions, when an engine is 
rusty and dirty on the outside, the inside 
of it will not bear careful inspection. 
The outward surfaces can be kept clean 
and bright when the machinery is in 
operation, consequently if a in an will 
not attend to this during working hours, 
he will probably not spend any more 
time taking care of the internal parts, 
while the machinery is shut down and 
other employes are enjoying themselves, 
than will barely keep them in operation. 

THE STRIKING POINTS 

To ascertain what may be called the 
"striking points" of an engine, disconnect 
the connecting rod from the crosshead, 
move the piston as far as it will go in one 
direction and put a mark on the guides 
at one end of the crosshead. Now move 
the piston as far as it will go in the op- 
posite direction and mark the guide in 
the same way. This reads very smooth- 
ly, and it will work in the same way, 
provided you take the packing rings out 
before the piston is moved further than 
it moves in regular service. If you fail 
to do this and one of them is caught in a 
port, it may cause more or less trouble 
to get it out. When the connecting rod 
and the packing rings have been put in 



368 STEAM ENGINEERING 

again, place the crank on each center 
alternately and mark the extreme travel 
of the crosshead as before. The differ- 
ence between these two marks at each 
end represents what is sometimes called 
the clearance. This is correct so far as 
the travel of the piston is concerned, but 
it does not include the waste space in 
ports, etc. An engine may be' so de- 
signed that it will wear unevenly, hence 
the piston may be drawn gradually 
towards one of the cylinder heads, as 
lost motion is taken up, but so long as 
the crosshead does not reach these marks 
on the guides, known as the "striking 
points," there is no danger of the piston 
striking either cylinder head. 

SMALL REPAIR JOBS 

Do not attempt to do a job of piping, 
or a small repair job on the engine during 
the noon hour, as it will probably take 
longer than you expect, hence when it is 
time to start the machinery for the after- 
noon you will not be ready, the other 
employes will be imable to work, and 
you will heartily wish that you had 
postponed it until evening, when there 
was plenty of time to do it without in- 
terfering with the regular work. 



STEAM ENGINEERING 



369 



N. 1-t I>. CO 00 CO 00 00 1>» - 

C0rJ<Tj<iO»C?OCDt^r^X_ 

C0C0fOCOCOC0C0COCOC0''i'Ti 



•"t*0»0«Ol>»t*"<d< 



^2 
OS 



0'OiO»C»OiOU5iOO«0'Om»CiC»OiOiOiO 



H r^ N r* rH lO O « r^ r-l r}< 

- . - -. J5a>00-<f-H(N(N(NCOCO 

<N IM (N (N (M (N (N (N (N CC CO CO CO CO CO CO CO CO 



I- 



r^ CO CO "<i< •<*' >o »o <o <o t^ c^ 00 00 05 o> 



Nk00>C<J»Ot>.O<Ni0t»O<N'#«000OC<)C0 
'-l--if-IC<IC<><NCOCOCOeO'*'*'*rrT}<iOiOiO 
MCqNMNW(N(NCaC<JNCvIoa(NC^(NCvlC<J 



o -^ M CO ^ »n o r^oo o> o '-' c<j CO Tj* lO «o r» 



s.,^ 



w 



^, 



ALPHABETICAL LIST OF TABLES 

Page Table 

Areas of boiler heads 73 16 

Areas of circles from 1 to 2 inches 36 5 

Areas of circles from 1 to 60 inches 239 42 

Areas of segments, determined by 

given f onniua 52 10 

Areas of segments, determined by 

ordinates ^ 61 9 

Average pressure of steam 272 45 

Boiler tubes from 1 to 4 inches 64 13 

Capacity of pumps in gallons 179 35 

Constants used in determining the area 

of segments 53 11 

Diameter and area of rivets 20 4 

Diameter and weight of water pipes .... 152 27 
Equivalent of common and decimal 

fractions 8 2 

Extra strong boiler tubes from J^ to 

4 inches 65 14 

Friction of water in iron pipes 194 37 

Fuel saved by feed water heaters 116 21 

Head and pressure of water 146 25 

Heating surface in boiler tubes 69 15 

Heating surface to develop one horse 

power.... 74 17 

Height of water on a piston to balance 

50 pounds steam pressure 187 36 

Horse power constants 282 46 

Horse power of high speed compound 

non-condensing engines 315 52 

Horse power of high speed compoxmd 

condensing engines, 14 expansions .... 333 55 
Horse power of high speed compound 

condensing engines, 16 expansions . . . 334 56 
Horse power of low speed compound 

non-condensing engines 312 51 

Horse power of low speed compound 

condensing engines, 14 expansions . . . 332 53 
Horse power of low speed compound 

condensing engines, 16 expansions . . . 333 54 



372 ALPHABETICAL LIST OF TABLES — Cont. 

Page Table 
Horse power of engines with 80 pounds 

pressure 291 48 

Hyperbolic logarithms 260 44 

Indicated horse power of high speed 

engines 295 50 

Indicated horse power of low speed 

engines 292 49 

Limit of suction power for cold water . . . 161 29 

Limit of suction power for hot water . . . 162 30 

Natural sines of angles 44 8 

Number of strokes required for a given 

piston speed 178 34 

, ressure and head of water 145 24 

roperties of water 92 18 

" « saturated steam 100 19 

Pounds of water per horse power 112 20 

Power required by single acting triplex 

pumps 141 22 

Power required by double acting triplex 

pumps 143 23 

Power required to raise water 149 26 

Ratios of expansion 256 43 

Safe loads for stay bolts and braces 37 6 

Safe pressures for steam boilers based 

on given rule 6 1 

Safe pressures for steam boilers based 

on U. S. Government rule 12 3 

Safe speed of fly wheels 289 47 

Size and capacity of single pumps 159 28 

Size and capacity of duplex pumps 171 31 

Size and capacity of compound pumps . . 175 32 

Size and capacity of pvmips and receivers 200 38 

Size and capacity of electric pumps .... 202 39 
Size and capacity of direct connected 

electric pumps 204 40 

Size of long stroke pumps 177 33 

Square inches supported by stay bolts 

and braces 40 7 

Standard boiler flues from 4J4 to 21 

inches 63 12 

Weight of water above 212° Fah 212 41 



NUMERICAL LIST OF TABLES 

Table Page 

1 Safe pressures for steam boilers based 

on given rule 6 

2 Equivalent of common and decimal 

fractions 8 

3 Safe pressures for steam boilers based 

on U. S. Government rule 12 

4 Diameter and area of rivets 20 

5 Areas of circles from 1 to 2 inches 36 

6 Safe loads for stay bolts and braces 37 

7 Square inches supported by stay bolts 

and braces 40 

8 Natural sines of angles 44 

9 Areas of segments, determined by 

ordinates 51 

10 Areas of segments, determined by 

given formula 52 

11 Constants used in determining the 

area of segments 53 

12 Standard boiler flues from 4^ to 21 

inches 63 

13 Boiler tubes from 1 to 4 inches 64 

14 Extra strong boiler tubes from H 

to 4 inches 65 

15 Heating surface in boiler tubes 69 

16 Areas of boiler heads 73 

17 Heating siurface to develop one horse 

power 74 

1 8 Properties of water 92 

19 a a saturated steam 100 

20 Pounds of water per horse power 112 

"21 Fuel saved by feed water heaters 116 

22 Power required by single acting triplex 

pumps 141 

23 Power required by double acting triplex 

pumps 143 

24 Pressure and head of water 145 

25 Head and pressure of water 146 

26 Power required to raise water 149 

27 Diameter and weight of water pipes , . . 152 

28 Size and capacity of single ptunps 159 

29 Limit of suction power for cold water. . . 161 



374 NUMERICAL LIST OF TABLES-Cont. 

Table Page 

30 Limit of suction power for hot water . . . 162 

31 Size and capacity of duplex pumps 171 

32 Size and capacity of compound pumps . . 175 

33 Size of long stroke pumps 177 

34 Number of strokes required for a given 

piston speed 178 

35 Capacity of ptimps in gallons 179 

36 Height of water on a piston to balance 

50 pounds steam pressure 187 

37 Friction of water in iron pipes 194 

38 Size and capacity of pumps and receivers 200 

39 Size and capacity of electric pumps .... 202 

40 Size and capacity of direct connected 

electric pumps 204 

41 Weight of water above 212° Fah 212 

42 Areas of circles from 1 to 60 inches 239 

43 Ratios of expansion 2^;^. 

44 Hyperbolic logarithms 260 

45 Average pressure of steam 272 

46 Horse power constants 282 

47 Safe speed of flywheels 289 

48 Horse power of engines with 80 pounds 

pressure 291 

49 Indicated horse power of low speed 

engines 292 

50 Indicated horse power of high speed 

engines 295 

51 Horse power of low speed compound 

non-condensing engines 312 

52 Horse power of high speed compound 

non-condensing engines 315 

53 Horse power of low speed compound 

condensing engines, 14 expansions . . . 332 

54 Horse power of low speed compound 

condensing engines, 16 expansions.. 333 

55 Horse power of high speed compound 

condensing engines, 14 expansions . . . 333 

56 Horse power of high speed compound 

condensing engines, 16 expansions . . . 334 



INDEX 
A 

Sec. Page 

1 Absolute pressure 1 97 

2 Accounting for moisture in steam . 1 106 

3 Actual horse power of boilers 1 76 

4 Angularity of braces 1 41 

5 Area of a segment of a circle 1 47 

6 Average and mean effective pres- 

sure 3 248 

1 Back pressure, and the ratio of 

expansion 3 254 

2 Benefits of feed water heaters .... 1 113 

3 Boiler, actual horse power of 1 76 

4 " and engine horse power 1 105 

5 " feeders 2 131 

6 " " receiver pumps as. . . 2 195 

7 " flues 1 59 

8 " head, rear, hand hole in the 4 347 

9 " plate, tensile strength of ... 1 2 

10 * shells, re-inforcing rings for . 1 4 

11 « test, object of 1 79 

12 " tubes 1 64 

13 Boilers, forcing 4 351 

14 " steam, safe working pres- 

sure of 1 5 

15 « steam, efficiency of 1 126 

16 " steam, heating surface of . 1 65 

17 " strength of 1 1 

18 Bracing fiat surfaces 1 32 

19 Braces, angularity of 1 41 

20 " surface supported by 1 38 

21 Bracing the heads of tubular 

boilers 1 45 

22 Brass and fibrous packed pistons . . 2 166 

23 Bright engines 4 366 



376 ' INDEX— Continued 



Sec. Page 

1 Capacity of injectors 2 215 

2 Circle, area of a segment of a 1 47 

3 Coal required to evaporate water. . 1 120 

4 Compound engines 3 3 

6 ■ " designing 3 3041 

6 • • mean effective 

pressure in ... . 3 309 

7 • " highspeed.... 3 314? 

8 * condensing engines, 3 

low speed 3 320 

9 * steam pumps 2 171 

10 Combustion, products of 4 352 : 

11 Comparison of thermometers ... . 1 89 

12 « "results 3 324 

13 Contents of water pipes 2 151 

14 Constants, horse power 3 280 

15 Consuming and preventing smoke. 4 352 

16 Crank pin heating 4 364 

17 " « boxes, filing 4 365 

18 Cylinders, steam and water 2 184 

19 • " diameter of 3 235 



1 Definitions 3 227 

2 Designing compound engines 3 304 

3 Determining the mean effective 

pressure 3 248 

4 Diameter of steam cylinders 3 235 

5 * « flues 1 60 

6 Direct acting steam pumps 2 153 

7 Directions for setting and operat- 

ing steam pumps 2 1 

8 Double strap, butt joint, strength 

of 1 

9 Draft, forced 4 3 

1 Duplex pumps 2 1 



INDEX — Continued 377 

E 

Sec. Page 

1 Effective pressure, mean, deter- 

mining the 3 248 

2 Effective pressure, mean, in com- 

pound engines 3 309 

3 Efficiency of steam boilers 1 126 

4 « « boiler feeders 2 3 

5 « ■ the injector 2 213 

6 Electric transmission of power, 

pumps driven by 2 200 

7 Engine and boiler, horse power ... 1 105 

8 Engines, bright 4 366 

9 " compound 3 302 

10 " " designing 3 304 

11 Engine, lining an 4 364 

12 " piston, pressure on 4 365 

13 ■ slide valve ..4 354 

14 " starting an 4 362 

15 ■ shutting down an 4 360 

16 " striking points of an 4 367 

17 « stop motion 4 362 

18 " selecting an 3 233 

19 Equivalent horse power 1 108 

20 • evaporation 1 117 

21 Evaporate water, coal required to . 1 120 

22 Evaporated water, per pound of 

combustible. 1 123 

23 ■ • from and at 

2120 Pah 1 125 

24 Expansion, ratio of, and the back 

pressure 3 248 



F 



1 Factor of safety 1 5 

2 Feeders, boiler 2 131 

3 Feed water heaters, benefits of 1 113 

4 Fibrous packed and brass pistons . 2 166 

6 Filing crank pin boxes 4 365 

6 Flat surfaces, bracing of 1 52 



378 



INDEX — Continued 



Sec. 

7 Flanging, special 1 

8 Flues, diameter of 1 

9 " safe pressure for 1 

10 " length of 1 

11 Fly wheels 3 

12 Forcing boilers 4 

13 Forced draft 4 

14 Fusible plug 4 



1 Gasket, rubber man hole 4 

2 Given head, pressure secured by a 2 

3 Governor, speed of 4 

H 



61 
286 
351 
353 
350 



349 
146 
363 



1 


Hand hole in rear boiler head 4 


347 


2 


Head necessary to give required 






pressure 2 


145 


3 


Heads of tubular boilers, bracing 






the 1 


45 


4 


Heating crank pin 4 


364 


5 


• surface of steam boilers. . . 1 


65 


6 


• * square feet of, per 






horse power 1 


74 


7 


Heaters, feed water, benefits of . . . 1 


113 


8 


High and low speed 3 


293 


9 


■ speed compound engines 3 


314 


10 


Horse power of boilers 1 


76 


11 


• ■ equivalent 1 


108 


12 


■ • required by pumps. . . 2 


139 


IS 


* ■ "to raise water 2 


147 


14 


• * of a steam engine .... 3 


238 


15 


* * constants 3 


280 


16 


■ ■ more about 3 


289 



INDEX — Continued 379 



Sec. Page 

1 Increase of power, providing f or . . 3 334 

2 Injector, efficiency of the 2 213 

3 Injectors 2 204 

4 « capacity of 2 215 

5 ■ types of 2 217 



1 Joint double strap butt 1 23 

2 ■ riveted 1 4 

3 ■ • strength of 1 14 



Length of flues 1 61 

Leveling tubes 4 351 

Lifting power of pumps 2 160 

Lining an engine 4 364 

Load and steam pressure 1 129 

Low and high speed 3 293 

• speed compound condensing 

engines 3 320 

M 

Man hole cover, putting on a 4 346 

" " gasket, rubber 4 349 

Mean effective pressure in com- 
pound engines 3 309 

Mean effective pressure, deter- 
mining the 3 248 

Mean effective and average pres- 
sure 3 248 

Moistvu-e in steam 1 83 

More about horse power 3 289 

N 

Necessary head to give required 



380 INDEX — Continued 

o 

Sec. 

1 Object of a boiler test 1 

2 Operating and setting steam pvimps 2 

3 Ordering steam ptunps 2 

4 Over and under, running 4 

P 

1 Piping injectors 2 

2 Pistons, brass and fibrous packed. . 2 

3 Plug, fusible 4 

4 Points, striking, of an engine 4 

5 Power pumps 2 

6 " required by pumps 2 

7 " "to raise water 2 

8 Practical points for progressive 

power plant operatives 4 346 

9 Pressure, absolute 1 97 

TO " mean effective, deter- 
mining the 3 248 

11 "on engine piston 4 365 

12 " required, necessary head 

to give 2 145 

13 " secured by a given head . 2 146 

14 Preventing and consuming smoke 4 352 

15 Products of combustion 4 353 

16 Providing for increase of power ... 3 334 

17 Pumps, compound 2 171 

18 " duplex 2 163 

19 " driven by electric trans- 

mission of power 2 200 

20 " lifting power of 2 160 

21 " receiver as boiler feeders . . 2 ' 195 

22 " setting and operating 2 189 

23 * single 2 155 

24 * steam, direct acting 2 153 

25 " steam, ordering 1 188 

26 " speed of 2 176 

27 Putting on a man hole cover 4 346 



INDEX — Continued 381 

Q 

Sec. Page 

Quality of steam produced 1 83 



Ratio of expansion, and back pres- 
sure 3 254 

Rear boiler head, hand hole in the 4 347 

Receiver ptimps as boiler feeders, 2 195 

Reinforcing rings for boiler shells, 1 4 
Required pressure, necessary head 

to give 2 145 

Results, comparison of 3 324 

Rivets, strength of 1 18 

Riveted joint 1 4 

• " strength of 1 14 

Rubber man hole gasket 4 349 

Running under and over 4 359 



Safety, factor of 1 5 

Safe loads for stay bolts and braces 1 37 

" pressure for flues 1 59 

" working pressure, U. S. rule 

for 1 10 

Saturated steam 1 96 

Segment of a circle 1 47 

Selecting an engine 3 233 

Setting and operating steam ptunps 2 189 

" the valves of duplex pvwips 2 165 

Shutting down an engine 4 360 

Single ptmips 2 155 

Slide valve engine 4 228 

Smoke, preventing and consuming 4 352 



382 INDEX — Continued 

s 

Sec. 

14 Special flanging 1 

15 Speed of governor 4 

16 Speed, low and high 3 

17 " of pumps 2 

18 Square feet of heating surface per 

horse power 1 

19 Starting an engine 4 

20 Steam boilers, efficiency of 1 

21 " • heating surface of. . 1 

22 * ■ safe working pres- 

sure of 1 

23 ■ cylinder, diameter of 3 

24 • engine, horse power of a ... . 3 

25 " engines, various types de- 

scribed 3 

26 " and water cylinders 2 

27 ■ pumps, compound 2 

28 « « direct acting 2 

29 • « ordering 2 

30 ■ pressure and load 1 

31 * produced, quality of ....... 1 83 

32 ■ velocity of 2 206 

33 Stop motion on an engine 4 362 

34 Strength of double strap butt joint 1 27 

35 Strength of plate 1 2 

36 « « rivets 1 18 

37 ■ * riveted joints 1 14 

38 " " steam boilers 1 1 

39 Striking points of an engine 4 367 

40 Surface supported by stay bolts and 

braces 1 38 



1 Tensile strength of boiler plate 1 

2 Thermometers 1 

3 Tubular boilers, bracing the heads 

of 1 



U^DEX— Continued 383 

T 

Sec. Page 

4 Tubes, boiler 1 64 

5 " leveling 4 351 

6 Types of injectors 2 217 

7 '* " steam engines described. 3 227 

U 

1 Under and over, running 4 359 

2 U. S. Government rule for safe 

pressures 1 10 

V 

1 Valves of duplex pumps 2 165 

2 - Various types of steam engines 

described 3 227 

3 Velocity of steam 2 206 

W 

1 Water and steam cylinders 2 184 

2 " evaporated under working 

conditions 1 104 

3 " evaporated from and at 

212" Fah 1 125 

4 '* evaporated per pound of 

combustible 1 123 

5 " horse power required to 

raise 2 147 

6 " pipes, contents of 2 116 



STEAM ENGINEERING 385 

COBBS 

High Pressure Spiral Piston and 
Valve Rod Packing 

Suitable for Steam, Gas and Air Pressure 
100 POUNDS AND OVER 




Spiral-Style 2 

IT is made with a rubber core, of an 
oil and heat resisting compound 
which is covered with a well frictioned 
duck and an outer covering of fine 
asbestos. It is guaranteed not to be- 
come hard, or score the rods. 

It has proven exceptionally service- 
able on the piston rods and valve stems 
of all kinds of engines, pumps, gas and 
air compressors, steam hammers, etc. 

"Cobbs" High Pressure Spiral Pack- 
ing is made both square and round, and 
is carried in stock in all sizes from yi in. 
to 13< in. 

Price per lb., $1.75 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



386 



STEAM ENGINEERING 



VULCAN 

High Pressure Spiral Piston and 

Valve Rod Packing 






i^s 




MSMLmMM 



Style 112 

TS made of a very fine quality of duck, 
well frictioned, with the thread end 
against the rod, making the best kind 
of wearing surface. It also has a rub- 
ber cushion which, when heated, will 
expand and force the duck against the 
rod, which has a tendency to steady it 
if it is a vibrating rod. The cushion is 
protected by a layer of duck which pre- 
vents it from sticking to the stuffing 
box. 

"Vulcan" High Pressure Packing is 
particularly serviceable on direct con- 
nected dynamo engines, high speed 
engines, rock drills, etc., particularly 
where there are vibrating rods. 

Carried in stock in all sizes from 5^ in. 
to 1 in. 

Price per lb., $1.75 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 387 

MAGIC 

Rubber Cushion Diagonal Expan- 
sion Spiral Packing 




Style 5 

IS especially adapted for service where 
the rod is scored. In cases where it 
is necessary to screw down tight on the 
glands, "Magic" Packing gives splendid 
results, set thumb tight. The wedge 
side has a soft rubber back which, when 
heated, expands, in some cases, suffici- 
ently to compensate for worn rods. 
The flax is braided and interwoven with 
a soft, frictionless metal wire, binding it 
together so there is no possible chance 
for the rod to pick up strands of the 
flax to get into cylinders and pass from 
there into the check valves. Magic 
Packing is filled with the best cold 
lubrication that can be procured. It 
gives best service for ammonia, steam, 
boiling water plungers, pistons, etc. 
Carried in stock in all sizes from % in. 
to 1 in.; larger sizes made to order. 

Price per lb., $1.50 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



388 STEAM ENGINEERING 

B. & P. 

H. P. Gum Cushion Asbestos Spiral 1j 
Piston and Valve Rod 
Packing 

Suitable for Steam, Gas and Air Pressure 

150 POUNDS AND OVER 




Style 1003 

/^UR High Pressure Cushion Asbestos 
Spiral and Ring Packings are made 
of the best quality Asbestos, with a soft 
red cushion. There is a growing de- 
mand for a Packing to withstand ex- 
tremely high pressures and superheat, 
which can be supplied with our High 
Pressure Cushion Asbestos, which is 
made in Spiral. Rod, Coil and Rings and 
carried in stock in sizes from % in. to 
l^in. 

Price per lb., $1.75 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 389 

GRAPHON 

Low Pressure Spiral Piston and 

Valve Rod Packing 

Suitable for Steam Pressure 

LESS THAN 100 POUNDS 




Style 1 

\X/E make for Steam, Ammonia, Hot 
or Cold Water, in Spiral, Plain and 
Sectional Cup Rings. We realize the ne- 
cessity of lubricating this style of Pack- 
ing for the particular uses to which it 
will be applied and after years of study 
and experiment, we have discovered a 
process of lubrication for steam and 
ammonia, which gives our Graphon 
Packing the enviable position of second 
to none. 

Carried in stock in sizes from Vs in. 
to 1 K in. 

Price per lb., $1.50 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



390 STEAM ENGINEERING 

HIGH PRESSURE 

Waterproof Hydraulic Packing 

Suitable for Cold Water Pressure 

UP TO 3,000 POUNDS 




T S strictly a Waterproof High Pressure 
Packing. It is thoroughly lubricated 
according to our secret process, which 
makes it impervious to water. The 
rubber core acts as a cushion. It is 
guaranteed to 3,000 pounds pressure 
and is most serviceable on High Pres- 
sure Pumps of all kinds and deep mine 
work. Made in Spiral, Coil and Rings 
and carried in stock from 14, in. to 1 5^ 

in. 

Price per lb., $1.75 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 



391 



COBBS High Pressure Ring Piston 

and Valve Rod Packing 
Suitable for Steam, Gas and Air Pressure 

100 POUNDS AND OVER 




Cobbs High Pressure Round Ring Packing 
Style 102 




Cobbs High Pressure Square Ring Packing 
Style 203 

•'/^OBBS" Piston and Valve Stem Packing 
\_^ is the outcome of years of experiments 
by its inventor, Mr. J. H. Cobb, who has 
been with the New York Belting & Packing Com- 
pany, for many years. 

See page 385 for Cobbs Spiral. 
Price per lb., $2.00 Write for Discount 
NEW YORK BELTING & PACKING CO. 



r 



392 STEAM ENGINEERING 

VULCAN 

High Pressure Square Ring 
Packing 




Style 212 

'T'HIS packing is manufactured of the 
same material and for the same pur- 
poses as "Vulcan" High Pressure Spiral 
Packing. 

When ordering give diameter of rod 
and inside diameter of stuffing box. 

See page 386 for Spiral style. 

Price per lb., $2.00 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 393 

MAGIC 

Diagonal Expansion Ring Packing 

Suitable for Hot Water 

100 TO 212 DEGREES 




Style 105 

COR Piston Rods and Boiler Feed 
and outside packed Plunger Pumps, 
where water is pumped from wells, 
traps, heaters. 

This packing is manufactured of the 
same material and for the same pur- 
poses as "Magic" Spiral Packing. 

When ordering give diameter of rod 

and inside diameter of stuffing box, also 

state whether to be used for water, 

steam or ammonia. 

See page 387 for Magic Wedge Spiral. 

Price per lb., $2.00 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



B. & P. 

High Pressure Cushion Asbestos 

Ring Piston and Valve Rod 

Packing 

Suitable for Steam, Gas and Air Pressure 

150 POUNDS AND OVER 




T^O-DAY, the tendency is toward ex- 
tremely high pressure and superheat. 
To overcome these severe conditions, 
we offer our B. & P. High Pressure As- 
bestos Ring Packing. 

See page 388 for H. P. Asbestos 
Spiral. 

Price per lb., $2.00 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 



395 



GRAPHON Low Pressure Plain and 

Sectional Rings 

Suitable for Steam Pressure 

LESS THAR 100 POUNDS 




Graphon Sectional Ring Packing Style 201 

In sending sizes of rods and stuffing boxes, 
please state if possible, the class of work for 
which they are intended, as we make consider- 
able difference in the lubrication when used for 
Steam and Ammonia. 

See page 389 for Graphon Spiral. 
Price per lb., $2.00 Write for Discount 
NEW YORK BELTING & PACKING CO. 



396 STEAM ENGINEERING 

HIGH PRESSURE 
Waterproof Hydraulic Ring 

Packing 

Suitable for Cold Water Pressure 

UP TO 3,000 POUNDS 




Style 206 

LJIGH PRESSURE Waterproof Hydrau- 
lic Packing is made from long fibre 
flax and treated with our special water- 
proof compound, which makes it imper- 
vious to water, and work with little or no 
friction. These Rings £re made for Cold 
Water Pistons, Plungers on Cold Water 
Pumps, Elevator Cylinders, Hydraulic 
Presses, Stern Glands and special hydrau- 
lic work. 

See page 390 for High Pressure Water- 
proof Hydraulic Spiral. 

Price per lb., $2.00 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 397 

COBBS Throttle Packing 
Suitable for Steam Pressure 

UP TO 250 POUNDS 

For use on Locomotive Throttles. 
(In sending orders always specify O. D. of the 
rod, I. D. of the stuffing box, and the depth of 
the box. There is no exact standard for the sizes 
of these Throttle Packing Sets as there is so much 
variation in the different locomotives.) 




Style 1820 

Does Not Get Hard 
Cannot Wear the Rod 

IT has a rubber core and a woven asbestos 
cover. It does not become hard in service as 
do other throttle packings, but remains soft and 
pliable, consequently it is very durable. 

There is nothing made that gives such good 
service. 

This packing is put up in sets as shown in cut 
and we strongly recommend that it be purchased 
in this form. We can, however, supply this 
packing in spiral coils if desired, put up in boxes 
of weights and lengths as shown on page 385. 

Price per set, $2.00 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



398 



STEAM ENGINEERING 



COBBS Air Pump Packing 
Suitable for Steam and Air Pressure j 

UP TO 250 POUNDS 

For use on New York and Westinghouse Air • 
Pumps and Air Compressors. 

(Put up in double boxes and exact fitting sets.) 

Note that this packing is put up in "double 
sets,"_ each tied up in rings and packed in boxes. 
This is sufficient Packing for both ends of "West- 
inghouse" Air Pump, although the "N. Y. Du- 
plex" Pumps require two (2) double sets. 




Style 1830 
Does Not Get Hard Cannot Wear the Rod 

Sizes of Air Pump Packing are as follows: 
WESTINGHOUSE AIR BRAKE PUMP 

Ins. Diam. Out. Diam. 
8 in. depth 4 rings . . 1 '/4 in. . . 2 in. 



11 i 



1 ' 



5 " . . IK in. . . 2|/^in. 

NEW YORK AIR BRAKE PUMP 
No. 2 Duplex 

Depth 4 rings. Ins. Diam. 1 '/4 in. Out. Diam. 2 in. 
No. 5 Duplex 

Depth 5 rings. Ins. Diam. !}< in. Out. Diam. 2'/4 in. 

It has a rubber core and a woven asbestos cover. 
It does not become hard in service as do other air 
pump packings, but remains soft and pliable, con- 
sequently it is very durable. 

There is nothing made that gives such good 
service. 

Price per set, $2.00 write for Discount 

NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 



399 



INDESTRUCTIBLE White Sheet 
Packing 




Style 10 

THIS is a Sheet Packing that will stand very 
high pressure and extreme heat. It is not 
affected by ammonia, liquor, or alkali, making 
a tight joint when used for steam, air, hot or cold 
water. It does not dry out or become hard and 
brittle, but retains its tough elastic qualities 
either in or out of service. It will not blow out 
under any pressure, and the joint can be broken 
numerous times without renewing the packing. 
Carried i stock in sizes ^7^2 in. to 14 in. 

Price per lb., $1.00 

Write for Discount 

Indestructible White Sheet Packing 

With Brass Wire Insertion 

Style 29 

For marine work, vibrating joints, or where 
there are but a few bolts in flange, or where the 
ports run close to the edge, and hard places to 
hold in general. 

Made M.6 and Vs in. thick. 

Price per lb., $1.25 

Write for Discount 
i NEW YORK BELTING & PACKING CO. 



400 STEAM ENGINEERING 

RUBY Red Sheet Packing 




Style 11 

THIS packing is made of a soft tenacious com- 
pound of red rubber that conforms to the in- 
equalities of flanges and will not blow out 
under the highest pressures. In the hottest joint it 
remains soft and flexible, and can be removed 
and used again and again in the same joint, 
packing it as perfectly as when first inserted. 

It also makes a perfect packing for hydraulic 
and cold joints of all kinds. It perfectly with- 
stands the action of steam, alkalies, ammonia, etc. 
The application of graphite to the face of the 
packing before use prevents adhesion to flanges 
and makes it easier to break the joint after use. 
Carried in stock from ys2 in. to 3^ in. and 36 in. 
wide. 

Price per lb., $1.00 Write for Discount 

Ruby Sheet Packing 
With Brass Wire Insertion 

Carried in stock \ie in. and J4 in. — 36 in. wide. 
Price per lb., $1.25 Write for Discount 
NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 401 

KARBONITE Black Sheet Packing 




Style 12 

THIS packing is made from the highest 
. grade of rubber combined with the 
Fvery best ingredients for the purpose, 
I producing a Packing that does not vul- 
jcanize. It will not harden when sub- 
I jected to extreme heat and will not blow 
I out under the highest pressure. The 
; action of oils, alkalies, ammonia, etc. 
is resisted better by "Karbonite" than 
by any other known rubber compound. 
j It will take up the contraction of metal 
when line cools off, and prevents joints 
from leaking. Nothing produced will 
iast as long or give such good service 
iLinder any and all conditions. 

Carried in stock from Vsz in. to ^ in. 
Price per lb., $1.25 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



402 STEAM ENGINEERING 

SALAMANDA Sheet Packing 




Style 13 

A HIGH-GRADE self-vulcanizing 
packing. Because of its semi-cured 
condition it readily conforms to rough 
surfaces, and when subjected to heat 
hardens, making a secure, permanently 
packed joint. 

It can be used between unfinished 
cast iron flanges, saving the cost of 
turning and planing. 

Carried in stock in thicknesses from 
Ysz in. upward — 36 in. wide. Wire 
insertion'made to order. 

Price per lb., 65 cents 

Write for Discount 
NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 403 

FIRO Superheat Sheet High Pres- 
sure Steam Packing 




Style 1846 



Has every qualification the name implies, and for 
Gas Engines, Automobiles, Motor Boats. Steel 
Plants, Turbines; in fact any place where extra- 
ordinary high pressures and intense heat exist, 
Firo has no equal. Firo will never vulcanize, 
cannot burn or char in a joint, will not flatten out 
between the flanges when the bolts are drawn up 
and makes a positively permanent joint, insuring 
economy to the plant and safety to the workmen. 
It is absolutely non-absorbent. Firo is made in 
sheets 40 in. x 40 in., 40 in. x 80 in. and 40 in. x 
120 in., in all thicknesses from ^^2 in. up. 

A sheet 40 in. x 40 in. will weigh as follows: 

ya2 in. thick 3 H pounds, 

He in. " 61^ " 

Hin. " 13 

Price per lb., $2.00 write for Discount 

ASBESTOS Metallic Sheet Packing 

Style 410 

Is made of the pure Asbestos Yarn, each strand 
having two brass wires interwoven, which give it 
strength. It is made 40 in. wide in all thicknesses 
from y32 in. up and fof all purposes where the 
usual fibrous and rubber sheet packings would be 
destroyed by intense heat. 

Price per lb., $1.50 write for Discount 

NEW YORK BELTING & PACKING CO. 



404 STEAM ENGINEERING 

CLOTH INSERTION 

Sheet Packing 




COR steam, hot or cold water, and 
weather strips. 
We make three grades of superior 
quality of this packing in the following 
brands : 



"Double Diamond" 


Style 33 


"Carbon" 


" 34 


"Spider" 


" 14 



In thicknesses of from V32 in. to ^ in. 
1-ply of duck to each Vie in. in thickness 

Write for price. 

WIRE INSERTION 
Sheet Packing 

Chiefly used in joints where C. I. 
Packing would be quickly burned out 
by heat or blown out by excessive 
pressure. 

NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 405 

RUBBER BELTING 




Test Special Rubber Belting 

Friction Surface 

A SUPERIOR belt for all power trans- 
mission purposes. — Main drives, line ' 
shaft, counter shaft and machine drives. 

Test Special Rubber Belting is univer- 
sally known for its strength and the ability 
to transmit every ounce of power gen- 
erated. Wasteful slippage is overcome by 
the positive hold given by the friction 
surface and the adhesion of the plies of 
duck is so strong that separation is 
avoided. 

Test Special Rubber Belting shows 
economy in initial cost, and its freedom 
from stretch is a protection against trouble 
in operation. 

Our Booklet "Test Special Rubber Belt- 
ing" should be in every Engineers library. 
Complete Engineering data. Horse Power 
Tables, Care of Belting, etc., contained in 
this handy volume will help in the solution 
of many belting problems. 
. Your copy will be sent on request. 

NEW YORK BELTING & PACKING CO. 



406 STEAM ENGINEERING 

STEAM REGULATOR 
DIAPHRAGMS 

N account of the severe work they are sub- 
jected to, we make them of a high grade 
rubber, most suitable for the purpose. 
Below, we show cuts and give sizes in which 



they are made. 



o 




High and low pressure. Made to withstand 
the most severe service. 

No. 1, 6% in. outside Diam., each • . . $ .55 
No. 2, 9 " " " ■ « _ .85 

No. 3 115^ " " " " .. 1.45 

No. 4, 15 " « " " .. 2.65 

PUMP DIAPHRAGMS 




For "Edson" and "Loud" pumps. These are 
high grade in every sense of the word. 

No. Edson, each $1-50 

No. 2 " " 2.00 

No. 3 « " .... 3.00 

No. 4 " « 4.00 

No. 1 Loud, " 2.00 

No. 2 « " 3.00 

Write for Discounts 
NEW YORK BELTING & PACKING CO. 



^ 



STEAM ENGINEERING 407 

PUMP VALVES 




"THE proper working of a Pump de- 
pends mostly on the Valves. It is 
necessary that these should be of the 
very best material and of a density 
suitable for the work the Pump must 
perform. 

Our experience in manufacturing 
Valves for Marine, Blower Engines, 
Acid, Mining, Boiler Feed and other 
Pumps, enables us to furnish the proper 
Valves and puts us in a position to 
make any special Valve you may desire. 
It is essential that you state explicitly 
the conditions under which the Valves 
are to be used, the size of the hole, the 
thickness and width of the Valves; also 
working pressure. 

NEW YORK BELTIXG & PACKING CO. 



408 STEAM ENGINEERING 

HARD RUBBER VALVE DISCS 




New Style Old Style 

For all Makes of Valves 

/WIANUFACTURED of hard rubber 
composition for the highest steam 
pressures, also ammonia, oils, acids, etc. 
Will not soften or wear uneven, thus 
preventing leaky valves; a rubber com- 
position that has been tested by many 
years of actual service and warranted 
to give satisfaction. 

Also furnished in soft composition 
for cold water, low pressures if desired. 

In ordering state whether New Style 
(oblong hole) or Old Style (round hole) 
is wanted. 

Carried in stock in all sizes as listed. 





List Prices 




Size 


Price Each 


Size 


Price Each 


>iinch 


$.03 


1 14 inch 


$.09 


H " 


.04 


13^ " 


.12 


H " 


.04 


2 


.18 


H " 


.05 


2H " 


.24 


1 " 


.06 


3 


.33 



NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 



409 



THE VULCANITE EMERY 
WHEEL 




IT can be safely said, to-day, that there is 
hardly a plant of any kind that does not use an 
Emery Wheel for some purpose; if for nothing 
else, to sharpen the tools of the workmen. Where 
Emery Wheels are used, the vexatious question 
arises, "Which is the safest Wheel to use?" 

The Vulcanite is universally conceded to be the 
strongest and best Emery Wheel made. 

By our process, the Emery is thoroughly and 
evenly mixed with best Para rubber, then forced 
into moulds and vulcanized under great pressure. 
The result is a Wheel of extraordinary strength 
and uniform consistency — one that is absolutely 
safe — will cut fast — stand up well on the corners 
and is easy to operate. 

VULCANITE EMERY WHEELS FOR 
ALL PURPOSES 

Malleable Iron Castings Grey Iron Castings 



Steel Castings 
Wrought Iron 
Machine Shop Work 
Stove Castings 
Car Wheels 
Dental Purposes 
Twist Drills 



Brass Castings 

Rough Work in General 

Plow Points 

Drop Forgings 

Car Couplings 

Tool Work 

Agricultural Implements 



Gumming and Sharpening Saws 
And many other purposes 

NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 

RUBBER HOSE 




MADE in our three well known 
'^* brands "1846 Para," "Double 
Diamond " and " Carbon," also " Inde- 
structible " brand. The latter is made 
with a special woven inner jacket, 
manufactured under our patent, and 
protected by rubber cover of high 
quality. 

We manufacture a complete line of 
Fire, Water, Steam, Brewers' Suction, 
Air Brake, Tender, Acid, Garden, Oil, 
Air drill, Pneumatic tool. Chemical and 
other hose, and are prepared to make 
any special hose for any specific pur- 
pose our trade may desire. 

NEW YORK BELTING & PACKING CO. 



STEAM ENGINEERING 411 

GOODS THAT WE 
MANUFACTURE 



Pneumatic Tool Hose 
Vacuum Cleaning Hose 
Water Hose 
Steam Hose 

Brewers' Hose 
Mill Hose 
Fire Hose 

Suction Hose 

Air Brake Hose 
Tubing (Pure, Machine and Cloth Inserted) 
Garden Hose 
Packings 

Gaskets and Rings 
Valves 

Dredging Sleeves 
Mats and Matting 

Rubber Belting 

Bradley Hammer Cushions 
Oil Well Packers 

Rubber Covered Rolls 

Fire Department Supplies 
Molded Goods 

Interlocking RubberTiling 
Etc., Etc 

NEW YORK BELTING & PACKING CO. 



J, 



1'^)^-^ 
^^^M 




: :»^I«f '^ 



LIBRARY OF CONGRESS 



021 213 192 1 



